STEAM    TURBINES 


PRACTICE  AND   THEORY 


BY 


LESTER  G.  FRENCH,  S.  B. 

yi 

Mechanical  Engineer 


FIRST  EDITION 


McGRAW-HILL  BOOK  COMPANY 

239  WEST  39TH  STREET,  NEW  YORK 
6  BOUVERIE  STREET,  LONDON,  E.  C. 

1907 


COPYRIGHT,  1907 

BY 
LESTER  G.  FRENCH 


PREFACE 


This  book  had  its  beginning  in  the  editorial  office  of  Machinery, 
New  York.  For  nine  years  previous  to  July,  1906,  the  author 
was  editor-in-chief  of  Machinery,  and  it  was  during  the  latter 
half  of  this  period  that  the  steam  turbine  became  a  commercial 
success  in  America  and  finally  reached  its  dominant  position  in 
the  field  of  electric  generating. 

Inasmuch  as  a  live  technical  journal  is  constantly  in  touch  with 
events  in  its  field,  the  author  had  an  exceptional  opportunity  to 
collect  notes  and  data  on  the  steam  turbine  and  allied  subjects 
during  the  important  formative  period  of  the  turbine,  when  it  was 
more  discussed,  probably,  than  any  engineering  subject.  Advan- 
tage was  taken  of  this  opportunity  to  begin  what  has  grown  to  be 
a  large  collection  of  such  notes  and  data,  which  have  served  as 
the  foundation  for  the  present  volume. 

The  author  is  responsible  for  various  articles  and  paragraphs 
which  have  appeared  in  Machinery  upon  the  subject  of  the  tur- 
bine, some  of  which  are  here  reproduced  in  more  or  less  modi- 
fied form.  Most  of  the  matter  in  these  pages,  however,  was  writ- 
ten especially  for  them  and  has  not  been  used  before. 

In  brief  explanation  of  the  contents,  it  may  be  said  that  the 
first  chapter  is  a  condensed  treatise  upon  the  fundamental  princi- 
ples of  the  steam  turbine.  The  second  chapter  traces  its  early 
development  and  shows  the  "state  of  the  art"  at  the  time  when  the 
turbine  became  a  commercial  success.  It  is  believed  that  this 
review  will  not  only  save  considerable  research,  but  in  connection 
with  the  first  chapter  will  serve  to  ground  the  beginner  in  first 
principles ;  for  there  is  no  way  of  accomplishing  this  so  effectively 
as  by  a  study  of  what  others  have  done. 

Much  attention  has  been  devoted  to  the  results  of  tests  upon 
the  flow  of  steam,  upon  the  action  of  steam  on  vanes  and  upon  the 
economic  performance  of  turbines.  For  comparison  with  the  lat- 
ter there  is  also  a  review  of  tests  upon  reciprocating  engines,  and 
enough  data  are  included  to  enable  an  intelligent  comparison  to 

333953 


iv  PREFACE 

be  made  under  different  running  conditions,  with  due  allowance 
for  the  efficiency  of  generators  and  engines. 

The  mathematical  treatment  has  been  limited  mainly  to  a  dis- 
cussion of  the  adiabatic  flow  of  steam  and  to  the  principles  of 
turbine  vanes,  which  together,  are  the  basis  of  all  turbine  calcula- 
tions. Preliminary  to  the  calculations  upon  the  flow  of  steam  is 
a  chapter  for  reference,  upon  the  properties  of  steam,  which  in- 
cludes, also,  an  explanation  of  the  temperature-entropy  diagram, 
and  data  upon  the  specific  heat  of  superheated  steam — subjects 
constantly  recurring  in  modern  writings  upon  heat  and  steam. 
The  attempt  has  been  to  simplify  the  mathematical  treatment  as 
much  as  possible,  and  illustrative  examples  have  been  worked  out 
wherever  it  was  thought  they  would  be  helpful. 

The  commercial  and  operative  sides  of  the  subject  have  re- 
ceived due  attention,  since  no  broad  grasp  of  any  important  en- 
gineering subject  can  be  had  by  viewing  it  solely  from  the  stand- 
point of  the  technicist.  Under  this  heading  may  also  be  classed 
the  treatment  that  is  given  of  high-vacuum  condensing  systems, 
which  it  is  generally  admitted  cause  engineers  more  trouble  than 
the  turbines  themselves. 

Other  reference  to  the  contents  seems  unnecessary  except  to 
say  that  the  descriptive  part  of  the  text  was  intended  to  be  compre- 
hensive, but  free  from  padding.  The  author  wishes  to  express 
his  appreciation  for  assistance  rendered  and  information  given  by 
many  engineers  and  friends ;  to  the  publication  departments  of  the 
several  turbine  manufacturing  companies  for  material  so  courte- 
ously supplied ;  and  to  technical  periodicals  from  which  informa- 
tion has  been  drawn  and  to  which  credit  has  been  given  at  the 
proper  places  throughout  the  text.  The  author  is  especially  in- 
debted to  The  Industrial  Press,  publishers  of  Machinery,  for  per- 
mission to  use  matter  and  engravings  that  have  appeared  in  the 
columns  of  that  journal. 

L.  G.  F. 

Brattleboro,  Ft.,  January,  1907. 


CONTENTS 


CHAPTER  i.     STEAM   TURBINE  PRINCIPLES 1 

Impulse  and  Reaction — Essential  Features  of  the  Turbine — How 
Steam  and  Water  Turbines  Differ — Difference  between  Turbines 
and  Piston  Engines — Steam  Nozzles — Distinction  between  Impulse 
and  Reaction  Turbines — Steam  Turbine  Types. 

CHAPTER  II.     EARLY  STEAM   TURBINE  PATENTS 22 

A  Review  of  the  Essential  Claims  of  Important  Turbine  Patents. 

CHAPTER  III.     SIMPLE  IMPULSE  TURBINES 67 

The  De  Laval  Steam  Turbine — Special  Applications  of  the  De 
Laval  Turbine. 

CHAPTER  IV.     THE  PELTON  AND  SIMILAR  TYPES 80 

Rateau's  Simple  Impulse  Wheel — Riedler-Stumpf  Turbine — Rich- 
ards' Design — The  Kerr  Turbine. 

CHAPTER  V.     COMPOUND  IMPULSE  TURBINES — MULTICELLULAR  TYPE...     95 
The  Rateatt  Turbine — The  Zoelly  Turbine — The  Hamilton-Holz- 
warth  Turbine. 

CHAPTER  VI.     COMPOUND  IMPULSE  TURBINES   (Continued) 113 

The  Curtis  Turbine — The  Riedler-Stumpf  Turbine. 

CHAPTER  VII.    REACTION    TURBINES 135 

Parsons  Turbines — The  Westinghouse-Parsons  Turbine — The 
Brown-Boveri  Turbine — The  Allis-Chalmers  Turbine. 

CHAPTER  VIII.     MISCELLANEOUS  TURBINES  AND  APPARATUS 159 

Combined  Impulse  and  Reaction  Turbines — The  British-Westing- 
house  Turbine — The  Sulzer  Brothers  Turbine — The  Landmark 
Turbine — The  Rateau  Steam  Accumulator  System. 

CHAPTER  IX.    STEAM     TURBINE     PERFORMANCE — COMPARISONS     WITH 

THE  STEAM  ENGINE 173 

Conversion  of  Power  Units — Efficiency  of  Engines  and  Generators 
— Calculations  involving  Efficiency — Thermal  Unit  Basis  of  Per- 
formance— Tables  of  Tests  on  Turbines — Comparison  of  Turbines 
and  Engines. 

CHAPTER  X.    STEAM  TURBINE  PERFORMANCE  (Continued) 198 

Characteristics  of  Turbines  under  Variable  Loads — Results  of 
Turbine  and  Engine  Tests  under  Variable  Loads — Effect  of 
Vacuum  upon  Economy — Effect  of  Superheating — Economy  with 
Change  in  Speed. 


vi.  CONTENTS 

CHAPTER  XI.    EXPERIMENTS  ON  THE  FLOW  OF  STEAM 217 

Napier's  Rules — Experiments  of  Brownlee,  Kunhardt,  Kneass, 
Rosenhain,  Rateau,  Gutermtith  and  Others. 

CHAPTER  XII.     STEAM  AND  ITS  PROPERTIES 247 

Notation  and  Definitions — Heat  Values  for  Steam  and  Water — 
Temperature-Entropy  Diagrams — Characteristic  Equations  for 
Adiabatic  Expansion — Specific  Heat  of  Superheated  Steam — Spe- 
cific Volume  of  Superheated  Steam. 

CHAPTER  XIII.    CALCULATIONS  ON  THE  FLOW  OF  STEAM 266 

Equations  for  Saturated  Steam — Calculations  upon  Superheated 
Steam — Steam  Nozzle  Design — Practical  Considerations. 

CHAPTER  XIV.    TURBINE  VANES 284 

The  Vanes  of  Impulse  Turbines — Calculations  of  Efficiency  and 
Elements  of  Velocity  Diagrams — Diagram  for  Pelton  Wheel — 
Diagrams  for  Compound  Impulse  Turbines — Diagrams  for  Re- 
action Turbines — Tests  on  Buckets  and  Channels. 

CHAPTER  XV.    BODIES  ROTATING  AT  HIGH  SPEED 305 

Action  of  High-speed  Bodies — Methods  of  Balancing — Stresses  in 
Rotating  Bodies. 

CHAPTER  XVI.    NOTES  ON  EFFICIENCY  AND  DESIGN 314 

Calculation  of  Efficiency — Losses  in  Turbines — Temperature-En- 
tropy Diagram  Applied  to  Stage  Turbines — Example  in  Design. 

CHAPTER  XVII.    THE  COMMERCIAL  ASPECT  OF  THE  TURBINE 327 

Relative  Advantages  of  Turbines  and  Engines — Relative  Space 
Occupied  by  Turbines  and  Engines — Space  Required  for  Condens- 
ing Apparatus — Comparative  Cost  of  Turbine  Outfits  and  Their 
Maintenance — Turbine  Troubles — Blade  Erosion. 

CHAPTER  XVIII.     CARE  AND  MANAGEMENT 358 

General  Directions — Operating  the  De  Laval  Turbine — Operating 
the  Parsons  Turbine — Operating  the  Curtis  Turbine. 

CHAPTER  XIX.    CONDENSING  APPARATUS  FOR  HIGH  VACUUM 371 

Gain  from  High  Vacuum — Surface  Condenser  Plants — Jet  and  In- 
jector Condensers — Data  upon  Condenser  Performance. 

CHAPTER  XX.    THE  STATUS  OF  THE  MARINE  TURBINE 393 

History — Atlantic  Liners  Fitted  with  Turbines — Turbine  Boats  of 
the  Cunard  Line — Comparison  between  Marine  Turbines  and  Re- 
ciprocating Engines. 


STEAM  TURBINES 


CHAPTER  I 

STEAM  TURBINE  PRINCIPLES. 

The  steam  turbine,  like  the  water  turbine,  is  based  on  the  prin- 
ciple that  when  a  fluid  is  in  motion  its  energy  will  be  con- 
verted into  mechanical  work,  if  the  fluid  impinges  on  moving 
vanes  which  change  its  direction  of  flow  and  reduce  its  velocity. 
It  differs  from  the  water  turbine  in  important  particulars,  how- 
ever, due  to  the  facts  that  water  and  steam  have  very  different 
properties  and  that  the  steam  turbine,  like  the  steam  engine,  is 
a  heat  motor  and  must  utilize  the  heat  energy  of  steam. 

The  Principle  of  the  Water  Turbine  is  illustrated  in  Fig.  1, 
which  shows  the  effect  of  a  curved  vane  upon  a  stream  of  water. 


Fig.  1.     Effects  of  Stationary  and  Moving  Vanes  upon  a 
Stream. 

The  lines  w,  R,  V ,  etc.,  represent  velocities  and  also  show  direc- 
tion of  motion.  At  A  the  vane  is  supposed  to  be  stationary 
and  the  stream  glides  upon  it  tangentially,  without  shock,. at  a 


2  STEAM  TURBINES 

velocity  R,  and  leaves  it  tangentially  at  the  same  velocity.  The 
only  effect  of  the  vane  is  to  alter  the  direction  of  flow. 

If  the  vane  now  be  given  a  velocity  w  in  the  direction  shown,  a 
particle  of  water  starting  at  a  will  reach. point  c  by  the  time  the 
tip  of  the  vane  has  traveled  from  a  to  b.  The  direction  and  ve- 
locity of  the  stream  relative  to  the  vane  will  then  be  represented 
by  the  line  be,  and  it  is  evident  that  the  water  will  meet  the  vane 
with  considerable  shock  and  in  this  instance  will  fail  to  touch  it 
anywhere  except  at  the  tip. 

To  bring  about  tangential  action  of  the  jet  while  the  vane  is 
moving,  which  is  essential  to  smooth  and  economical  running,* 
either  the  nozzle  must  be  given  a  motion  w,  or  else  the  direction 
and  velocity  of  the  stream  must  be  changed  to  V ' ,  the  resultant 
of  R  and  w.  (Shown  at  B.)  The  motion  of  the  water  relative 
to  the  moving  vane  will  then  be  the  same  as  its  motion  relative 
to  the  stationary  vane  before  the  change  was  made.  Similarly, 
in  leaving  the  vane  the  motion  r,  relative  to  the  vane  is  tangential 
and  equal  to  R.  But  as  the  stream  and  vane  each  has  the 
velocity  w  at  the  exit,  the  absolute  or  real  velocity  of  the  stream 
will  be  the  resultant  of  r  and  w,  or  z/.f  The  difference  between 
V  and  v  represents  that  part  of  the  stream  velocity  which,  neg- 
lecting losses,  has  been  transformed  into  wheel  velocity  w.  We 
thus  have : 


*This  is  the  theoretical  statement  of  the  conditions. 

fThe  subject  of  relative  motion  can  be  made  clear  by  the  simple  device  herewith. 
Two  pins,  a  and  b,  are  driven  in  a  board  and  a  cord  is  stretched  between  them.  Place 
an  oblong  piece  of  paper  under  the  cord  in  the  position  shown  and,  holding  it  sta- 


tionary with  one  hand,  draw  a  diagonal  line  across  the  paper  by  following  the  cord  with 
the  point  of  a  pencil.  Remove  the  paper  and  c  d  (at  the  right)  will  be  the  line  drawn. 
Again  place  the  paper  in  its  former  position  and  run  the  pencil  along  the  cord,  at  the 
same  time  pulling  the  paper  to  the  right  just  fast  enough  so  that  when  the  pencil 
reaches  the  opposite  edge  of  the  paper  point  c  will  come  opposite  arrow  d  marked  on 
the  board.  The  line  traced  this  time  will  be  c  k.  In  both  cases  the  pencil  followed 
the  same  course,  relative  to  the  board,  but  its  motion  relative  to  the  paper  was  not  the 
same  when  the  paper  moved.  In  like  manner  the  motion,  relative  to  the  vane,  of  the 
jet  of  water  in  Fig.  1  is  different,  when  the  vane  moves,  from  what  it  is  when  the  vane 
is  stationary. 


STEAM  TURBINE  PRINCIPLES  3 

Energy  given  to  the  moving  vane  by  each  pound  of  water  = 


In  this  case,  therefore,  the  effect  of  the  vane  is  not  only  to  change 
the  direction  of  the  stream,  but  to  reduce  its  velocity,  by  which 
means  its  kinetic  energy  is  changed  into  mechanical  work  at  the 
turbine  wheel. 

Impulse  and  Reaction. 

Impulse  and  Reaction  of  a  Jet. — The  dynamic  pressure  upon 
the  vane  of  a  turbine,  which  causes  the  rotation  of  the  wheel,  is 
the  result  of  the  impulse  and  reaction  of  the  impinging  jet.  Ac- 
cording to  older  writers  in  mechanics,  an  impulse  is  a  force  act- 
ing in  a  forward  direction  and  a  reaction  is  the  equal  and  opposite 
force — a  force  acting  in  a  backward  direction  relative  to  the 
impulse.  These  are  the  commonly  accepted  meanings,  although 
strictly  an  impulse  is  not.  a  force  in  the  sense  of  being  a 
push  or  a  pull,  but  is  a  term  used  to  express  the  same  meaning 
as  momentum,  when  a  force  acts  for  a  very  short  time,  as  when 
one  body  gives  another  a  sharp  blow.  A  reaction  is  properly  de- 
fined as  a  force. 


Fig.  2.     Impulse. 


Fig.  3.     Reaction. 


In  Fig.  2  is  a  tank  from  which  water  issues  through  a  nozzle 
and  impinges  against  a  flat  surface  capable  of  moving  horizontally 
— in  this  case  the  face  of  a  plank  suspended  so  that  it  is  free  to 


4  STEAM  TURBINES 

swing.  When  the  water  strikes  the  plank  the  latter  will  swing 
to  the  right  under  the  pressure  due  to  the  impulse  of  the  jet. 
As  the  jet  leaves  the  nozzle,  however,  it  exerts  a  reaction  against 
the  tank,  which  is  equal  and  opposite  to  the  force  due  to  the  im- 
pulse of  the  jet.  This  may  be  illustrated  by  suspending  the  tank, 
as  in  Fig.  3,  so  that  it  is  free  to  swing,  when  it  will  move  to  the 
left,  owing  to  the  reaction  of  the  jet. 


Fig.  4.     Measuring  Pressure  due  to  Impulse. 

The  explanation  of  the  reaction  is  that  there  is  no  pressure 
on  that  part  of  the  tank  where  the  nozzle  is  situated  and  the  un- 
balanced pressure  on  that  part  directly  opposite  to  the  nozzle  will 
therefore  tend  to  move  the  tank  in  a  direction  opposite  to  that 
in  which  the  water  is  escaping.  The  faster  the  water  escapes,  the 
greater  the  pressure  required  in  the  tank  to  give  the  water  its 
velocity  and  hence  the  greater  the  unbalanced  pressure  which  we 
call  the  reaction.* 

Impulse  and  Reaction  Upon  Curved  Vanes. — In  Fig.  2  the  ve- 
locity of  the  jet  is  suddenly  checked  when  it  strikes  the  plank, 
and  there  is  more  or  less  commotion  of  the  particles  of  the  fluid, 
causing  loss  of  energy.  This  is  always  the  case  when  there  is 
impact  of  the  particles  against  a  surface  and  is  to  be  avoided  by 
using  a  curved  surface  so  placed  as  to  change  the  direction  of 
flow  gradually  without  shock  or  jar,  as  previously  explained. 

*The  reaction  is  not  determined  solely  by  the  pressure  in  the  tank.  Since  action  and 
reaction  are  always  equal,  the  reaction  upon  the  tank  must  be  that  due  to  the  momen- 
tum of  the  jet.  The  velocity  and  weight  of  fluid  discharged,  however,  and  hence  the 
momentum,  depend  in  part  upon  the  shape  of  the  nozzle,  and  its  effective  area,  which 
is  never  exactly  equal  to  the  measured  area,  owing  to  the  coefficient  of  contraction  of 
the  jet. 


STEAM  TURBINE  PRINCIPLES  5 

Such  an  arrangement  is  shown  in  Fig.  4,  where  a  nozzle  N  is  dis- 
charging against  a  plate  D  attached  to  the  rod  R.  The  rod  is 
supported  by  guides  and  together  with  the  plate  can  move  longi- 
tudinally. The  pressure  of  the  jet  against  the  plate  is  balanced  by 
the  weight  W  supported  by  a  cord  passing  around  the  pulley  P 
and  attached  to  the  rod  R.  The  plate  D  is  so  shaped  at  the  center 
as  to  gradually  guide  the  particles  of  water  through  an  angle  of 
90  degrees,  thus  avoiding  the  impact  present  in  Fig.  2,  and  the 
particles  leave  the  plate  in  a  direction  parallel  with  its  face.  It  is 
evident  that  the  pressure  against  the  plate  is  due  solely  to  the 
impulse  of  the  jet  and  that  the  reaction  of  the  water  in  leaving 
the  plate  has  no  tendency  to  move  the  plate  longitudinally. 

In  Fig.  5  the  case  is  different.  Here  the  water  strikes  a  curved 
surface  and  is  turned  back  upon  itself  through  an  angle  of  180 
degrees.  This  surface  is  therefore  acted  upon  by  two  forces,  both 


Fig.  5.       Measuring  the  Combined  Effect  of  Impulse  and  Reaction. 

tending  to  move  it  to  the  right.  The  first  is  that  due  to  the  impulse 
of  the  jet,  just  as  in  Fig.  4,  which  acts  until  the  central  point  C 
of  the  curved  surface  is  reached ;  and  the  second  is  the  reaction 
of  the  jet,  which  begins  where  the  jet  starts  to  flow  backward 
and  continues  up  to  the  edge  where  it  discharges.  These  forces 
would  be  equal  if  there  were  no  frictional  or  other  losses,  and 
it  would  require  a  weight,  W,  just  twice  as  heavy  as  the  weight 
in  the  first  example  to  balance  the  end  thrust. 


6  STEAM  TURBINES 

Essential  Features  of  the  Turbine. 

How  Steam  and  Water  Turbines  Differ. — In  a  steam  turbine 
the  energy  of  the  steam,  due  to  its  motion,  is  converted  into  me- 
chanical work  at  the  turbine  wheel  by  the  use  of  curved  vanes,  as 
outlined  for  water  turbines.  There  are  two  important  distinc- 
tions between  steam  and  water  turbines,  however,  which  it  will 
be  advisable  to  refer  to  here.  These  are : 

First,  provision  must  be  made  in  the  steam  turbine  for  con- 
verting the  heat  energy  of  the  steam  into  kinetic  energy,  or  the 
energy  of  motion.  To  accomplish  this  the  passages  and  nozzles 
of  the  steam  turbine  must  be  designed  to  control  the  expansion  of 
the  steam  in  a  way  to  augment  its  velocity.  In  the  De  Laval 
turbine  the  expansion  is  effected  in  an  "expanding  nozzle"  which 
directs  the  steam  jet  against  the  blades  of  the  wheel.  The  walls  of 
the  nozzle  diverge  in  the  direction  of  flow  of  the  steam  so  that 
its  outlet  area  is  larger  than  its  inlet  area,  whereas,  in  a  water 
nozzle  the  outlet  area  would  be  smaller  than  the  inlet  area. 

Second,  the  steam  turbine  must  be  adapted  to  the  high  veloc- 
ities of  steam,  which  have  no  parallel  in  hydraulic  work,  although 
velocities  of  over  300  feet  a  second  are  met  with  in  water  power 
plants  on  the  Pacific  slope.  The  enormously  high  velocity  with 
which  steam  flows  from  an  orifice  often  exceeds  the  speed  of  a 
rifle  bullet.  The  new  Springfield  rifle  adopted  by  the  United 
States  Army  gives  an  initial  velocity  to  its  bullet  of  2,300  feet 
per  second,  or  over  26  miles  a  minute,  and  this  is  almost  exactly 
the  velocity  with  which  steam  at  50  pounds  gauge  pressure  would 
issue  from  a  nozzle  of  the  best  shape  when  discharging  into  the 
atmosphere.  In  single-wheel  impulse  turbines,  operating  under 
higher  pressures  and  with  a  condenser,  velocities  of  3,000  to  4,000 
feet  per  second,  or  even  more,  are  attained.  The  problem  of  deal- 
ing with  a  fluid  capable  of  attaining  such  incomprehensible 
velocities  calls  for  serious  consideration.  By  some  means  or 
other  the  rotating  elements  must  be  kept  within  the  speed  of 
safety  to  guard  against  rupture  of  material,  cutting  of  bearings, 
etc.,  and  at  the  point  where  power  is  to  be  used  the  speed  must  be 
within  the  practical  limits  of  convenience. 

The  Successful  Steam  Turbine. — It  will  be  evident  from  the 


STEAM  TURBINE  PRINCIPLES  7 

foregoing  that  the  successful  steam  turbine  must  accomplish  the 
three  following  results : 

First,  as  much  of  the  heat  energy  of  the  steam  as  possible  must 
be  converted  into  kinetic  energy. 

Second,  the  wheel  must  be  capable  of  utilizing  the  kinetic 
energy  of  the  steam  in  an  efficient  manner. 

Third,  the  apparatus  must  run  at  a  moderate  speed  at  the  point 
where  it  delivers  its  power,  and  all  parts  must  be  kept  within  the 
speed  of  safety. 

The  Steam  Turbine  Compared  ivith  the  Steam  Engine. — The 
fundamental  principle  in  any  economical  steam  motor,  whether 
turbine  or  piston  engine,  is  that  the  expansive  force  of  the  steam 
must  be  utilized.  The  direct-acting  steam  pump,  although  serving 
a  very  useful  purpose,  is  one  of  the  most  wasteful  steam  users  in 
existence.  Steam  is  supplied  to  the  cylinder  at  boiler  pressure 
throughout  the  whole  stroke.  It  forces  the  piston  ahead  because 
of  the  static  pressure  back  of  it  in  the  boiler,  in  the  same  manner 
that  water  would  do  if  the  steam  cylinder  of  the  pump  were  con- 
nected to  a  city  water  main.  Such  expansion  as  occurs  takes 
place  in  the  boiler  at  constant  temperature. 

Steam,  however,  is  capable  of  much  better  use  than,  this,  be- 
cause it  is  supplied  with  a  store  of  heat  energy  which  has  the 
power  to  make  the  steam  in  the  cylinder  expand  and  push  the 
piston  forward,  after  all  communication  with  the  boiler  has  been 
cut  off.  In  an  engine  working  expansively  steam  is  admitted  at 
boiler  pressure  until  the  point  of  cut-off  is  reached.  Up  to  this 
event  the  action  is  the  same  as  in  the  steam  pump,  but  during  the 
rest  of  the  stroke  the  piston  is  pushed  ahead  as  a  result  of  the  heat 
energy  of  the  steam  encased  in  the  cylinder. 

In  the  steam  turbine  the  process  of  the  expansion  engine  is  du- 
plicated, except  that  the  flow  of  the  steam  is  continuous  instead 
of  intermittent.  It  was  explained  that  steam  is  first  forced  into 
the  engine  cylinder  by  the  pressure  in  the  boiler  and  then  is  al- 
lowed to  expand  in  virtue  of  its  own  internal  heat  energy.  In 
the  turbine  the  steam  is  continuously  pushed  into  the  nozzle  by 
the  higher  pressure  at  the  inlet,  and  during  the  passage  through 
the  nozzle  it  expands  continuously  because  of  its  internal  energy. 
Each  particle,  as  it  expands,  pushes  the  particles  ahead  of  it  for- 


8  STEAM  TURBINES 

ward  at  a  faster  rate  and  so  increases  the  velocity  of  flow. 
Although  the  turbine  and  piston  engines  are  different  in  outward 
form,  they  are  equivalent  in  the  thermodynamic  action. 

The  difference  in  the  form  of  the  turbine  and  engine  is  due 
to  the  fact  that  the  turbine  is  designed  to  operate  by  changing 
the  motion  of  flowing  steam,  on  the  principle  of  the  water  turbine, 
while  the  engine  is  designed  to  operate  by  the  direct  pressure  of 
the  steam.  The  turbine  is  a  velocity  motor  and  the  steam  engine 
a  pressure  motor.* 

Steam  Nozzles. 

The  Study  of  Steam  Nozzles  of  Great  Importance. — Since  a 
turbine  operates  by  changing  the  motion  of  flowing  steam,  atten- 
tion must  be  given  to  the  proportions  of  the  passages  through 
which  the  steam  flows.  In  the  De  Laval  and  some  other  turbines, 
the  steam  flows  through  nozzles  which  direct  it  against  the  blades 
of  the  rotating  wheel.  In  other  machines  it  flows  through  pas- 
sages between  guide  vanes  which  form  what  is  virtually  a  group 
of  nozzles.  In  still  others  of  the  Parsons  type  both  the  stationary 
guide  vanes  and  the  blades  of  the  wheel  have  the  same  functions 
that  a  collection  of  nozzles  placed  side  by  side  would  have. 
Whatever  the  arrangement  of  the  passages  of  a  turbine,  through 
which  the  steam  passes,  they  may  be  regarded  as  steam  nozzles, 
provided  the  steam  fills  them  completely,  leaving  no  air  spaces, 
just  as  water  fills  completely  all  the  space  in  a  water  nozzle 
used  in  connection  with  a  hose  pipe.  In  order  to  understand  the 
action  of  steam  flowing  through  the  passages  of  a  turbine,  it  is 
necessary  to  study  its  action  in  flowing  through  nozzles  of  dif- 
ferent shapes,  and  the  principles  discovered  may  then  be  applied  in 
proportioning  the  ducts  or  passages  of  the  turbine. 

*It  must  not  be  imagined  that  the  driving  force  in  a  turbine  is  the  statical  steam 
pressure.  All  turbines  derive  their  power  from  changes  in  the  motion  of  the  working 
fluid.  The  pressure  in  a  turbine  might  be  infinite,  yet  unless  the  steam  possessed  the 
requisite  velocity  the  turbine  would  not  act.  On  the  other  hand,  if  the  steam  possessed 
the  requisite  velocity  then  the  pressure  might  be — if  that  were  practicable — absolutely 
zero,  and  yet  the  turbine  would  work  quite  normally.  The  statical  steam  pressure  acts 
equally  on  the  back  and  front  of  each  vane,  and  hence  produces  neither  end  thrust 
nor  rotation.  As  the  steam  passes  through  the  turbine  the  direction  of  its  motion  is 
altered  by  the  vanes,  and  hence  these  vanes  must  exert  force  on  the  steam.  It  is  this 
force,  or,  rather,  the  corresponding  reaction  of  the  steam  on  the  vanes,  which  causes  the 
rotation  and  where,  as  in  the  Parsons  turbine,  the  vanes  are  not  symmetrical,  an  end 
thrust. — From  "The  Theory  of  Steam  Turbines,"  by  Frank  Foster,  In  The  Engineering 
Review,  London,  May,  1904. 


STEAM  TURBINE  PRINCIPLES  9 

Flow  of  Steam  Through  Nozzles. — The  illustration,  Fig.  6, 
shows  the  effects  of  nozzles  of  different  shapes  upon  steam  flowing 
through  them.  The  general  form  of  the  jets  issuing  from  the 
nozzles  is  substantially  as  shown  by  Strickland  L.  Kneass,  engi- 


EXPANDS  TO  ATMOSPHERIC  PRESSURE  HERE 


Fig. 


6.     Types  of  Steam  Nozzles  and  the  Shapes  of  Jets  Dis- 
charging from  them. 


10  STEAM  TURBINES 

neer  of  the  injector   department,   William   Sellers   &  Co.,  inc., 
Philadelphia,  in  his  ''Practice  and  Theory  of  the  Injector." 

In  Fig.  6,  five  styles  of  mouthpieces  are  illustrated,  and  with  the 
exception  of  No.  5,  steam  is  supposed  to  be  flowing  from  some 
higher  pressure  down  to  atmospheric  pressure.  In  No.  5,  steam  is 
assumed  to  discharge  into  a  closed  chamber  in  which  the  pressure 
is  maintained  at  the  "critical"  pressure  referred  to  below.  The  five 
styles  of  mouthpieces  are  as  follows : 

1.  Short  cylindrical  tube,  slightly  rounded  inlet. 

2.  Nozzle  with  converging  walls — such  as  would  be  used    to 
produce  a  solid  water  jet. 

3.  Diverging  nozzle   with   rounded   inlet   and   straight  taper 
sides. 

4.  Diverging  nozzle,  rounded  inlet  and  diverging  sides,  curved 
as  shown. 

5.  Same  as  No.  1. 

6.  Orifice  in  a  thin  plate. 

While  the  nozzles  shown  are  supposed  to  be  of  a  circular  cross 
section,  it  is  evident  that  it  is  the  area  of  the  section  which  is  of 
importance  and  not  the  shape  of  the  section.  A  rectangular- 
shaped  section  would  answer  as  well  as  a  circular  section,  for  all 
practical  purposes,  and  in  fact  is  oftener  used  for  the  steam  pas- 
sages between  the  vanes  of  turbines. 

A  peculiarity  of  the  flow  of  steam  through  nozzles  is  that  the 
absolute  pressure  in  the  throat  of  the  nozzle  does  not  vary  greatly 
from  58  per  cent  of  the  absolute  initial  pressure,  unless  the  abso- 
lute pressure  at  discharge  should  be  more  than  58  per  cent  of 
the  initial  pressure.  It  can  be  shown  that  theoretically  this  should 
be  the  case,  although  it  is  known  from  tests  that  the  pressure  may 
be  somewhat  more  or  less  than  this.  The  pressure  of  58  per  cent  of 
the  initial  pressure  is  called  the  critical  pressure  and  has  an  impor- 
tant bearing  upon  the  subject  of  the  flow  of  steam. 

Nozzles  with  Parallel  Walls  or  Converging  Walls. — Many  ex- 
periments have  been  conducted  on  the  flow  of  steam  through 
nozzles  like  Nos.  1  and  2  and  it  is  found  that  when  discharging 
into  a  medium  which  has  not  more  than  58  per  cent  of  the  initial 
pressure  (equal  to  the  critical  pressure)  the  velocity  of  discharge  is 
not  far  from  the  constant  value  of  1,450  feet  per  second  and  the 


STEAM  TURBINE  PRINCIPLES  11 

weight  discharged  is  also  a  constant  quantity.  Suppose  that  one 
of  these  nozzles  were  arranged  to  discharge  into  a  tank,  the  pres- 
sure in  which  could  be  controlled  by  a  valve,  and  that  at  the  start 
the  tank  pressure  was  almost  equal  to  the  initial  pressure,  but 
was  gradually  decreased  until  it  finally  became  zero.  We  should 
find  that  the  velocity  and  weight  of  discharge  from  the  nozzle 
gradually  increased,  as  the  difference  in  pressure  increased,  until 
the  point  was  reached  where  the  tank  pressure  became  58  per  cent 
of  the  initial  pressure.  At  that  instant  the  velocity  of  discharge 
would  be  between  1,400  and  1,500  feet  per  second  and  any  further 
lowering  of  the  pressure  in  the  tank  would  have  no  effect  on  the 
velocity  or  weight  discharged.  The  discharge  would  remain  at  the 
same  rate  per  second,  however  low  the  final  pressure  became,  even 
if  carried  down  to  vacuum. 

In  nozzles  like  Nos.  1  and  2  the  pressure  of  the  steam  does  not 
drop  much,  if  any,  below  the  throat  pressure  until  the  outlet  is 
Beached,  but  when  the  steam  issues  from  the  nozzle  its  pressure 
suddenly  falls  to  that  of  the  medium  outside.  If  discharging  into 
the  atmosphere  it  would  fall  in  pounds  an  amount  equal  to  42  per 
cent  of  the  initial  pressure.  In  making  this  sudden  drop  the  steam 
is  free  to  expand  in  all  directions,  the  jet  enlarges  after  issuing, 
becomes  broken  up  and  scattered  and  is  inefficient  for  turbine  pro- 
pulsion. In  this,  as  in  all  cases  where  the  velocity  of  a  jet  is 
checked,  its  kinetic  energy  is  converted  back  into  heat,  generated 
by  the  friction  and  eddying  of  the  steam,  which  tends  to  superheat 
the  steam  at  discharge. 

Nozzles  for  Complete  Expansion. — Where  there  are  great  dif- 
ferences of  pressure,  a  more  efficient  steam  jet  can  be  secured  by 
using  a  diverging  nozzle,  like  No.  3,  such  as  has  been  employed 
for  many  years  in  steam  injectors.  In  such  a  nozzle  the  steam 
expands  to  the  lower  pressure  within  the  nozzle  itself  and  when 
the  steam  discharges  it  takes  the  form  of  a  solid  cylindrical  jet, 
equal  in  diameter  to  the  outlet  diameter  of  the  nozzle.  A  diverg- 
ing, or  more  properly  a  converging-diverging  nozzle,  has  a  con- 
verging inlet,  like  nozzle  No.  2,  to  which  is  added  a  diverging 
outlet,  for  the  purpose  of  controlling  the  expansion  of  the  steam 
beyond  the  throat  of  the  nozzle.  The  pressure  at  the  throat  is 
about  the  same  as  the  pressure  in  nozzles  1  and  2,  or  58  per  cent  of 


12  STEAM  TURBINES 

the  initial  pressure.  Beyond  the  throat  the  cross-sectional  area  in- 
creases just  sufficiently  to  accommodate  the  rapidly  increasing 
specific  volume  of  the  steam  (space  occupied  by  unit  weight) 
which  occurs  as  the  pressure  drops.  When  thus  proportioned, 
the  walls  restrict  the  expansion  of  the  steam  in  a  lateral  direction, 
but  allow  free  expansion  longitudinally.  Such  a  nozzle,  is  known 
as  an  expansion  nozzle  and  by  its  use  the  full  expansive  force  of 
the  steam  is  utilized  and  a  very  high  velocity  of  outflow  attained. 
A  nozzle  like  No.  4  has  been  found  to  give  slightly  better  results 
than  one  with  straight,  conical  sides,  like  No.  3,  but  it  is  more 
difficult  to  construct. 

Complete  expansion  may  be  obtained  in  a  straight  or  converging 
nozzle  by  arranging  so  that  the  pressure  of  the  medium  into  which 
it  discharges  shall  not  be  less  than  58  per  cent  of  the  higher  pres- 
sure. Under  these  conditions  the  steam  will  issue  from  the  nozzle 
in  straight,  parallel  lines,  or  nearly  so,  since  there  is  no  excess 
internal  pressure  to  make  the  jet  bulge,  and  all  the  expansive 
force  of  the  steam  will  be  expended  in  giving  velocity  to  the  jet. 
In  the  illustration,  nozzle  No.  5  is  intended  to  show  that  complete 
expansion  may  be  realized  in  a  straight  nozzle  discharging  into 
a  tank  in  which  the  pressure  is  58  per  cent  of  the  higher  pressure. 

Orifice  in  Thin  Plate. — When  steam  issues  from  an  orifice  in 
a  thin  plate,  as  in  nozzle  No.  6,  the  swelling  of  the  jet  after 
leaving  the  opening  is  even  more  marked  than  in  the  first  two 
nozzles  shown.  This  is  because  the  internal  pressure  of  the  steam 
is  higher  in  the  orifice  in  the  plate  than  at  the  mouth  of  the  tubes 
in  the  first  two  examples.  The  steam  has  an  opportunity  to  ex- 
pand more  fully  before  leaving  the  tubes  than  it  does  in  passing 
through  an  orifice  in  a  thin  plate. 

Distinction  Between  Impulse  and  Reaction  Turbines. 

Impulse  Steam  Turbine. — Both  water  and  steam  turbines  are 
grouped  into  two  general  classes  known  as  impulse  and  reaction 
turbines,  or  better,  into  action  and  reaction  turbines.  These 
terms  are  somewhat  misleading,  however,  because  all  practical 
turbines  operate  both  by  the  action  and  reaction  of  the  working 
fluid  and  it  would  be  clearer  to  designate  the  two  types  in  some 
other  way. 


STEAM  TURBINE  PRINCIPLES 


13 


In  Fig.  7  is  a  simple  impulse  turbine  having  curved  vanes 
against  which  the  jet  of  steam  impinges.  The  expansion  of  the 
steam  is  completed  within  the  nozzle,  and  there  is  no  expansion 
in  passing  through  the  wheel  passages.  The  pressure  between 
the  vanes  is  the  same  as  the  pressure  within  the  casing  in  which 
the  wheel  runs  and  the  steam  flows  freely  through  the  wheel 
passages  in  virtue  of  the  kinetic  energy  given  it  in  the  nozzle.  The 
wheel  is  driven  ahead,  first  by  the  pressure  due  to  the  impulse  of 
the  steam  and  then,  after  the  vanes  have  reversed  the  direction  of 
flow,  by  the  reaction  of  the  steam. 


Fig.  7. 


Usual  Type  of  Impulse 
Wheel. 


Fig.  8.     Impulse  Wheel  in  which  there 
is  no  Reaction. 


In  Fig.  7,  N  is  the  nozzle,  which  may  or  may  not  be  a  diverging 
nozzle,  according  to  the  pressure  against  which  it  is  discharging, 
and  W  is  the  wheel.  With  the  vanes  constructed  as  shown,  there 
would  be  spaces  S  S  not  filled  by  the  steam,  since  the  area  of  the 
passages  at  these  points  is  greater  than  at  the  entrance  and  exit. 
Some  manufacturers,  how  ever,  make  the  blades  thicker  at  the  center 
than  near  the  edges,  to  maintain  a  constant  area  and  so  avoid  pos- 
sible eddy  currents.  Although  a  so-called  impulse  wheel,  it  will 
be  evident  that  this  wheel  acts  both  by  impulse  and  reaction.  The 
chief  characteristic  of  this  type  is  that  the  expansion  occurs  wholly 
in  the  nozzle  or  guide  passages,  as  the  case  may  be. 


14 


STEAM  TURBINES 


What  would  strictly  be  an  impulse  wheel  is  shown  in  Fig.  8, 
where  the  vanes  are  so  curved  that  with  the  wheel  held  stationary 
the  steam  would  leave  them  in  a  direction  parallel  with  the  shaft. 
The  wheel  is  therefore  propelled  solely  by  the  pressure  against  the 
vanes  due  to  the  impulse  of  the  steam,  and  would  be  inefficient 
because  the  steam  would  have  a  high  residual  velocity  when  it 
left  the  wheel.  Such  a  wheel  is  on  the  principle  of  the  stationary 
vane  in  Fig.  4,  against  which  the  water  exerted  only  half  the  pres- 
sure that  it  did  when  the  force  of  reaction  was  taken  advantage  of. 

Reaction  Steam  Turbines. — In  Fig.  9  is  the  simplest  type  of 


Fig.    9. 


Two  forms  of  Reaction  Wheel. 


Fig.    10. 


reaction  wheel.  Here  the  steam  enters  the  trunnion  T,  flowing 
radially  outward  through  the  two  hollow  arms  A  A,  until  it  dis- 
charges through  the  nozzles  N  N.  The  arms  therefore  rotate  in 
a  direction  opposite  to  that  in  which  the  steam  escapes,  and  are 
driven  entirely  by  the  reaction  of  the  steam.  The  chief  difficulty 
of  this  type  of  wheel  is  the  excessively  high  speed  of  rotation. 
Supposing  the  wheel  to  be  perfectly  free  to  move,  its  momentum 
would  be  equal  to  the  momentum  of  the  escaping  steam.  The 
velocity  of  the  arms  would  theoretically  be  less  than  the  velocity 
of  the  steam  only  in  so  far  as  the  mass  of  the  arms  was  greater 
than  that  of  the  steam. 


STEAM  TURBINE  PRINCIPLES  15 

In  Fig.  10  is  shown  a  more  practical  form  of  reaction  wheel. 
Here  there  is  first  the  impact  of  the  steam  against  the  buckets, 
but  the  expansion  in  the  nozzle  is  only  partial  and  the  steam 
expands  still  more  and  acquires  additional  velocity  in  flowing 
through  the  wheel,  provision  being  made  for  this,  if  necessary, 
by  having  the  passages  diverge  in  the  direction  of  the  flow,  as 
shown  at  V.  The  steam  therefore  reacts  upon  the  wheel  when  it 
leaves  the  vanes  as  a  result  of  the  energy  acquired  in  the  wheel 
itself,  and  this  feature  gives  it  the  name  of  a  reaction  wheel.  It 
will  be  seen,  however,  that  the  wheel  acts  both' by  the  impulse  and 
the  reaction  of  the  steam  just  as  in  the  case  of  the  impulse 
wheel.  The  distinction  between  the  two  is  that  in  the  impulse 
wheel  the  expansion  of  the  steam  is  complete  within  the  nozzle 
and  in  the  reaction  ivheel  it  is  not  completed  until  after  it  enters 
the  zvheel  passage.  If  it  were  possible  to  attach  a  steam  gauge  to 
one  of  the  spaces  between  two  wheel  vanes,  it  would  show  a 
pressure  equal  to  the  pressure  of  the  medium  in  which  the  wheel 
was  turning  in  the  case  of  the  impulse  turbine  and  a  pressure 
higher  than  that  of  the  surrounding  medium  in  the  case  of  the 
reaction  turbine. 

Shape  of  Vanes  in  Impulse  and  Reaction  Turbines. —  The  illus- 
trations,* Figs.  11  and  12,  represent  the  guide  vanes  and  moving 


Fig.  11.     Shape  of  Vanes  in  Impulse  Wheels. 

vanes  of  impulse  and  reaction  turbines  respectively.  In  impulse 
turbines  the  wheel  vanes  are  symmetrical  or  nearly  so.  In  Fig. 
11,  G  G  are  the  guide  vanes  which,  if  the  fall  of  pressure  is  small, 
need  not  provide  diverging  passages.  The  wheel  vanes  at  V^ 
have  both  faces  parallel,  and  the  vanes  at  V2  are  thicker  at  the 

*From    a    paper    by    M.    J.    Rey    read    before    the    Societie    de    Ingeneuirs  Civils  de 
France,  March,  1904. 


16  STEAM  TURBINES 

center  than  at  the  edges,  forming  passages  of  a  uniform  width, 
as  explained  in  connection  with  Fig.  7. 


Fig.   12.     Shape  of  Vanes  in   Reaction  Wheels. 

In  Fig.  12  G  G  are  the  guide  vanes  and  V  V  the  wheel  vanes 
showing  in  a  general  way  the  contour  that  must  be  obtained  in 
reaction  turbines. 

Steam  Turbine  Types. 

The  Simple  Impulse  Turbine. — The  simplest  possible  arrange- 
ment of  the  steam  turbine  is  shown  in  the  diagramatic  sketch, 
Fig.  13,  where  N  is  a  nozzle  directing  a  jet  of  steam  against  the 
vanes  of  a  single  wheel  W  inclosed  in  a  casing.  This  is  like  the 
De  Laval  turbine,  which  must  utilize  steam  flowing  with  a  velocity 
of  3,000  to  4,000  feet  a  second,  and  the  peripheral  velocity  of  the 
wheel  should  be  nearly  one-half  of  this  to  utilize  the  total  energy 
of  the  steam.  In  the  De  Laval  turbine  the  peripheral  velocity  is 
frequently  as  high  as  1,200  feet  a  second  or  as  high  as  safety  will 
permit  with  the  strongest  materials  for  the  rotating  member. 
Such  high  velocity  of  rotation  makes  it  necessary  to  use  speed- 
reducing  gearing. 

Principle  of  the  Compound  Turbine. — In  turbines  of  large 
size  it  is  desirable,  and  in  fact  necessary,  to  avoid  such  high  speeds 
of  rotation  and  to  do  away  with  the  reducing  gears,  and  this 
is  accomplished  in  several  other  types  of  turbines  through  com- 
pounding. A  compound  turbine  may  be  built  either  for  water 
or  steam,  and  it  is  entirely  possible  for  a  water  jet  to  flow  at  such 
great  velocity  as  to  make  compounding  desirable  for  a  water 
turbine.  The  principle  of  compounding  is  very  simple  and  is  thus 
explained  in  Bodmer's  text  book  "Hydraulic  Motors" : 

"If  a  turbine  is  allowed  to  run  at  a  much  lower  speed  than  at 


STEAM  TURBINE  PRINCIPLES 


17 


the  best,  the  water  leaves  the  buckets  with  a  very  considerable 
absolute  velocity,  and  there  is  consequent  loss  from  unutilized 
energy.  This  energy  might,  however,  be  usefully  employed  in 
driving  a  second  turbine,  the  water,  after  leaving  the  first,  being 
deflected  by  a  set  of  stationary  guide  vanes  to  cause  it  to  enter  the 


4-W 


S        ri 


Fig.    13.     Simple    Im- 
pulse— De     Laval 
Type. 


Fig.    14.     Compound    Impulse 
— Riedler-Stumpf  Type. 

Steam  Turbine  Types. 


Fig. 


15.     Compound    Impulse 
— Curtis  Type. 


second  wheel  at  the  proper  angle.  Both  turbines  could  be  keyed 
to  the  same  shaft  and  their  speed  would  be  much  lower  than  that 
of  a  single  turbine  driven  by  the  same  head  of  water  and  utilizing 
it  to  the  same  extent.  This  arrangement  would  constitute  a  com- 
pound turbine  and  it  is  clear  that,  instead  of  two  wheels  only, 
three  or  more  might  be  employed  in  the  same  way,  the  speed  being 


18  STEAM  TURBINES 

lower  the  greater  the  number  of  wheels.  The  only  object  in  using 
a  compound  turbine  in  preference  to  a  single  one  would  be  to 
reduce  the  speed  in  cases  where  the  head  was  great  and  high 
velocity  of  rotation  inconvenient  or  impracticable." 

Compound  Impulse  Turbines. — There  are  different  methods  of 
taking  advantage  of  this  principle  of  compounding,  the  simplest 
one  of  which  is  shown  in  Fig.  14.  Here  steam  flows  through  the 
expansion  nozzle  N  which  reduces  its  pressure  to  that  of  the 
medium  in  which  the  turbine  wheels  rotate.  The  steam  then 
impinges  against  the  vanes  of  the  wheel  W ±.  Its  direction  is 
then  reversed  by  the  guide  vanes  G  and  it  next  impinges  against 
the  vanes  of  wheel  Wz,  on  the  same  shaft  S  as  the  other  wheel. 
In  this  example  the  action  is  the  same  as  outlined  above  by  Bod- 
mer,  and  the  steam,  having  acquired  its  velocity  in  the  nozzle, 
flows  through  the  passages  of  the  turbine  in  virtue  of  its  inertia. 
This  plan  has  been  carried  out  in  the  turbines  of  Professors 
Riedler  and  Stumpf,  although  the  arrangement  of  the  parts  is 
different. 

At  the  bottom  of  the  engravings  Figs.  13  to  17,  are  diagrams, 
the  upper  ones  of  which  show  the  change  in  velocity  and  the 
lower  ones  the  change  in  pressure  of  the  steam  in  the  different 
steps  of  its  progress.  In  the  first  illustration  the  velocity  of  the 
steam  increases  from  a  to  b,  and  as  it  flows  through  the  nozzle 
most  of  this  velocity  is  absorbed  by  the  wheel,  the  height  of  the 
line  c  d  indicating  the  residual  or  unused  velocity  of  the  steam  as 
it  leaves  the  wheel.  The  lines  1-2  and  2-3  show  that  the  pressure 
drops  in  the  nozzle  but  does  not  change  after  the  steam  strikes 
the  wheel. 

In  the  second  illustration,  Fig.  14,  the  velocity  increases  in  the 
nozzle,  but  in  the  guide  passages  where  no  work  is  done  it  re- 
mains constant,  or  would  do  so  except  for  frictional  losses,  and 
finally  in  the  last  wheel  the  velocity  drops  to  the  point  e.  The 
pressures,  however,  indicated  by  1  2-3  are  as  in  the  previous  case. 

Proportions  of  Passages  in  a  Compound  Turbine. — It  will  be 
noticed  that  the  height  x  of  the  passage  through  the  first  wheel 
Wlt  in  Fig.  14,  is  the  same  as  the  diameter  of  the  nozzle,  while 
the  height  of  the  passages  in  the  guide  G  and  the  second  wheel 
W2  is  slightly  greater.  The  drawing  is  made  in  this  way  to 


STEAM  TURBINE  PRINCIPLES  19 

emphasize  one  of  the  principles  of  the  flow  of  fluids  through  a 
turbine.  In  flowing  through  wheel  IV  lt  part  of  the  steam's  velocity 
is  given  up  to  the  wheel  so  that  the  velocity  in  guide  G  is  less 
and  more  room  must  be  allowed  for  the  slower  moving  fluids. 
No  increase  in  the  width  of  the  passage  in  wheel  IV lt  is  required, 
because,  while  the  absolute  velocity  of  the  steam  decreases  as  it 
progresses  through  the  wheel,  the  velocity  relative  to  the  moving 
vanes  does  not  change,  neglecting  frictional  losses.  For  the  same 
reason,  also,  the  passages  through  W z  are  the  same  in  height  as 
the  guide  passages  G.  As  a  matter  of  fact  a  turbine,  equipped 
with  nozzles  like  Fig.  14,  would  not  require  the  passages  to  be 
proportioned  as  above  outlined,  because  the  stream  of  fluid  would 
have  ample  room  by  spreading  out  around  the  periphery  as  it 
approached  the  exhaust  end. 

Modified  Form  of  Compound  Impulse  Turbine. — Difficulty  is 
experienced  in  making  steam  flow  through  groups  of  irregular 
passages  without  any  impelling  force  to  overcome  friction  other 
than  its  own  inertia.  To  supply  the  necessary  impelling  force  to 
make  up  for  frictional  losses  the  modification  shown  in  Fig.  15 
has  been  used,  notably  in  the  Curtis  turbine.  The  steam  as  before 
flows  through  diverging  passages  N  N  and  impinges  against  the 
vanes  of  the  first  wheel  and  then  passes  through  the  guide  vanes 
G  to  the  second  wheel.  The  depth  of  these  passages  increases 
slightly  from  the  point  where  the  steam  strikes  the  first  wheel  to 
the  point  where  it  leaves  the  last  wheel  at  g.  A  part  of  the  ex- 
pansion, however,  is  reserved  to  take  place  in  the  wheel  and  guide 
vanes  for  the  purpose  of  accelerating  the  velocity  sufficiently  to 
compensate  for  its  retardation  by  friction,  although  most  of  its 
expansion  occurs  in  the  nozzles  as  before.  This  turbine,  like  the 
two  previous  ones,  is  essentially  an  impulse  turbine,  though  the 
pressure  in  the  wheel  passages  is  slightly  higher  than  that  of  the 
medium  in  which  the  wheels  rotate. 

Stage  Turbines. — The  arrangement  of  Fig.  15  is  sometimes 
modified  by  having  two  or  more  groups  of  wheels  and  guides, 
each  of  which  is  in  a  separate  compartment  so  that  the  reduction 
in  velocity  and  pressure  is  not  as  great  in  any  one  step.  This 
plan  is  followed  in  the  Curtis  turbine,  where  each  compartment 


20 


STEAM  TURBINES 


Fig.     16.     Compound     Impulse  —  Multi-  Fig.    17. 

cellular  or  Rateau  Type. 

Steam  Turbine  Types. 


Compound    Reaction  —  Par- 
sons  Type. 


in  which  the  pressure  is  stepped  down  is  called  a  stage,  and  it  is 
known  as  a  multi-stage  turbine. 

The  simplest  form  of  stage  turbine  is  shown  in  Fig.  16,  which 
illustrates  the  principle  of  turbines  like  the  Rateau  and  Zoelly. 
This  is  in  effect  a  series  of  simple  turbines  like  that  of  Fig.  13, 
each  of  which  is  in  a  separate  compartment.  There  are  usually 
enough  compartments  so  that  the  steam  does  not  have  to  drop 
more  than  .4  of  its  initial  pressure  when  flowing  from  one  com- 
partment to  the  next,  and  diverging  nozzles  are  not  required. 
Steam  enters  through  the  guide  passages  A^  and  expands  down 


STEAM  TURBINE  PRINCIPLES  21 

to  the  pressure  in  the  first  compartment.  The  pressure  within 
the  wheel  passages  is  the  same  as  without  and  to  insure  a  uniform 
pressure  in  all  parts  of  the  compartment  holes  H  are  sometimes 
made  through  the  wheel  disks.  The  second  set  of  guide  pas- 
sages N2  has  a  larger  area  than  the  first  to  accommodate  the 
increased  volume  of  the  steam  and  the  same  is  true  of  the  suc- 
ceeding passages.  The  velocity  diagram  below  shows  that  there 
is  a  small  unused  residual  velocity  in  each  case  and  the  pressure 
diagram  shows  how  the  pressures  are  gradually  stepped  down. 

Compound  Reaction  Turbines. — We  now  come  to  the  reaction 
turbine  of  which  the  Parsons  turbine  is  the  most  prominent  rep- 
resentative. In  this  type  the  steam  expands  continuously  from 
boiler  pressure  to  vacuum.  There  are  alternate  rows  of  guides 
and  vanes,  the  latter  being  attached  to  the  drum  on  the  turbine 
shaft.  The  steam  flows  through  a  fixed  ring  of  directing  blades 
N  onto  a  revolving  ring  of  similar  blades  W  and  so  on,  its  pres- 
sure being  reduced  a  few  pounds,  say  two  or  three,  at  each  step. 
The  steam  finally  discharges  at  E.  It  will  be  seen  that  if  the 
wheel  were  fixed  and  the  steam  allowed  to  flow  through  the 
turbine  the  passages  themselves  taken  together  would  constitute 
a  large  expansion  nozzle,  and  the  flow  of  the  steam  would  in- 
crease from  beginning  to  end  as  shown  in  the  upper  diagram 
placed  beneath  the  section  of  the  wheel. 

Let  the  wheel  rotate,  however,  and  the  velocity  acquired  in 
passing  through  the  first  guide  ring  would  be  partially  absorbed 
by  the  first  wheel;  and  the  velocity  acquired  in  the  next  ring  of 
guide  blades  would  be  partially  absorbed  in  the  second  wheel, 
and  so  on.  The  line  of  velocities,  therefore,  would  be  represented 
by  a  b  c  d,  etc.,  in  the  middle  diagram.  The  pressure,  however, 
drops  gradually  from  beginning  to  end  as  represented  in  the  last 
diagram. 

This  diagram  shows  that  the  pressure  in  the  passages  of  the 
turbine  is  maintained  higher  than  the  final  pressure,  which,  as  has 
been  explained,  is  the  characteristic  of  the  reaction  principle. 

In  subsequent  chapters  modifications  of  these  simple  types  wilt 
be  shown,  but  what  has  been  given  is  believed  to  be  sufficient  to 
enable  the  reader  to  understand  the  descriptions  of  the  various 
turbines  which  follow. 


CHAPTER  11 

EARLY  STEAM  TURBINE  PATENTS. 

/ 

It/is  probable  that  the  first  steam  engine  was  a  turbine.  In 
Heroes  "Spiritalia,"  a  book  on  pneumatics  issued  in  the  second  or 
third  century,  is  a  description  of  the  whirling  eolipile  consisting 
of  a  small  hollow  sphere  mounted  on  trunnions,  one  of  which  is  hol- 
low for  the  admission  of  steam.  The  sphere  is  caused  to  rotate  by 
the  reaction  of  steam  flowing  from  two  diametrically  opposite  noz- 
zles having  bent  mouthpieces.  This  is  frequently  spoken  of  as  the 
beginning  of  the  reaction  turbine;  and  to  Branca,  who  issued  a 
work,  entitled  "The  Machine,"  published  at  Rome  in  1629,  is  given 
credit  for  the  first  impulse  wheel.  This  volume  contains  an  illus- 
tration of  an  eolipile,  in  the  form  of  a  negro's  head,  placed  over  a 
fire.  A  blast  of  steam  proceeds  from  the  mouth  and  impinges 
against  the  blades  of  a  large  wheel  which  it  was  proposed  to  con- 
nect by  means  of  cog  wheels  with  a  crude  stamping  mill  for  pul- 
verizing drugs.  These  very  early  efforts  could  have  been  nothing 
more  than  visionary  schemes,  but  they  are  scarcely  less  imprac- 
ticable than  many  of  the  later  inventions  to  be  found  in  the  pages 
of  the  patent  records.  Comparatively  few  of  the  steam  turbine  in- 
ventions embody  even  the  first  elements  of  success,  probably  be- 
cause most  of  those  who  have  directed  their  attention  to  the  subject 
have  failed  to  understand  either  what  was  required  or  what  means 
must  be  taken  to  accomplish  good  results. 

In  selecting  from  among  the  great  number  of  turbine  patents 
those  that  appear  to  have  useful  features,  the  author  has  had  in 
mind  the  requirements  of  the  successful  steam  turbine  as  outlined 
in  the  first  chapter,  and  has  not  given  space  to  inventions  unless 
they  seemed  to  embody  at  least  one  feature  that  would  contribute 
toward  a  practical  and  operative  machine.  With  the  exception  of 

*In  writing  this  review  the  author  has  drawn  on  the  historical  material  in  the 
valuable  series  "Roues  et  Turbines  a  Vapeur,"  by  M.  Sosnowki,  published  in  the  August, 
September,  October  and  November,  1896,  numbers  of  the  "Bulletin  de  la  Societe 
d'Encouragement  pour  1'Industrie  Nationale,"  Paris.  He  has  also  been  materially 
assisted  in  his  search  of  the  patent  records  by  the  list  of  English  turbine  patents  in 
Neilson*s  treatise,  "The  Steam  Turbine";  and  by  a  similar  list  of  United  States  patents 
kindly  supplied  by  Mr.  Robert  A.^  McKee,  mechanical  engineer,  steam  turbine  depart- 
ment, Allis-Chalmers  Company. 


EARLY  TURBINE  PATENTS 


23 


a  very  few  patents  taken  from  the  French  patent  records,  specifica- 
tions of  the  inventions  mentioned  are  to  be  found  either  in  the 
English  or  in  the  United  States  patent  records. 

Real  and  Pichon,  1827. — This  machine  operates  by  impulse  and 
is  one  of  the  earliest  attempts  to  produce  a  wheel  to  run  at  mod- 
erate speed  and  at  the  same  time  utilize  a  large  percentage  of  the 
energy  of  the  steam  by  the  principle  of  compounding.  Certain  of 


Fig.   1.     Real  and   Pichon   Compound  Turbine. 

the  details  of  the  original  patent  drawing  are  somewhat  obscure, 
but  the  illustration  has  been  made  to  correspond  with  the  text  as 
nearly  as  possible.  The  cylinder  A  contains  a  succession  of 
disks,  B,  which  divide  the  cylinder  into  compartments.  The  shaft 
F  is  turned  with  a  series  of  steps  upon  each  of  which  is  carried  a 
turbine  wheel  G,  having  short  radial  blades,  H,  around  its  pe- 
riphery. Steam  is  admitted  from  the  boiler  through  the  pipe  /  at 


24  STEAM  TURBINES 

the  top  into  the  first  compartment  and  flows  in  the  form -of  jets 
through  a  series  of  openings,  C,  against  the  blades  of  the  first 
wheel  which  runs  in  the  second  compartment.  The  steam  next 
passes  through  a  second  series  of  holes  in  the  second  disk  and  im- 
pinges against  the  second  wheel  and  so  on  to  the  bottom  of  the 
cylinder,  where  the  steam  exhausts  through  the  pipe  M.  The 
shaft  and  wheels  are  carried  by  a  step  bearing  and  power  is  sup- 
posed to  be  transmitted  through  the  gears  P  and  Q.  The  open- 
ings, C,  in  the  circumference  of  the  disks,  B,  are  bored  obliquely, 
so  the  steam  will  impinge  as  directly  as  possible  against  the  faces 
of  the  blades.  With  this  plan  the  pressure  will  drop  only  a  few 
pounds  from  chamber  to  chamber,  giving  the  steam  a  compara- 
tively low  velocity  of  flow. 

Avery  Turbine,  1831. — The  first  steam  turbine  patent  to  be  is- 
sued in  the  United  States  was  to  Foster  &  Avery  for  a  reaction 
wheel  of  the  Hero  type.  Strangely  enough,  this  is  one  of  the  few 
turbine  inventions  that  has  been  developed  and  put  into  actual  use, 
and  probably  it  is  the  only  steam  turbine  used  in  commercial  work 


-50 


Fig.   2.     Avery  Reaction  Wheel. 

in  this  country  until  a  considerably  later  date.  There  were  several 
of  these  machines  in  operation  in  1835,  some  of  which  were  used  to 
drive  saw  mills  near  Syracuse,  N.  Y.  In  1901,  Prof.  John  E.  Sweet 
contributed  a  description  to  the  Transactions  of  the  American  So- 
ciety of  Mechanical  Engineers,  accompanying  it  by  a  sketch  made 
from  an  original  drawing  of  the  Avery  wheel,  reproduced  in 
Fig.  2. 


EARLY  TURBINE  PATENTS  25 

The  arm  is  made,  with  the  exception  of  the  end  pieces  and  knife 
blades,  of  two  pieces  of  iron  brazed  together  from  end  to  end  at 
the  edges.  The  openings  at  the  ends  of  the  arms  for  the  steam  jets 
were  l/%  by  l/$  inch.  The  speed  of  the  tips  of  the  arms  was,  of 
course,  enormous.  Mr.  Avery  states  in  his  notebook  that  the  speed 
of  the  arms  of  a  7-foot  wheel  placed  upon  a  locomotive  in  1836, 
which  was  put  upon  a  railroad  near  Newark,  N.  J.,  and  ended  its 
life  in  a  ditch,  was  at  one  time  14^2  miles  a  minute  at  the  pe- 
riphery. A  difficulty  met  with  was  the  end  pressure  on  the  hollow 
shaft,  which  was  overcome  by  running  the  end  of  the  shaft  against 
the  edge  of  a  wheel  set  at  right  angles.  The  trouble  in  setting  up. 
the  packing  around  the  hollow  shaft  became  a  serious  matter.  It 
was  also  found  that  the  knife  edges  at  the  end  of  each  arm  were 
cut  away  by  the  steam  and  required  frequent  renewal.  The  noise 
also  was  very  objectionable.* 

Leroy,  1838. — In  commenting  on  Avery's  invention  Prof.  John 
E.  Sweet  has  said  that  he  long  had  the. conviction  that  expanding 
nozzles  applied  to  the  Avery  turbine,  in  place  of  the  plain  orifices 
used,  would  give  the  benefit  of  expansion  and  produce  superior 
results. 


*In  a  paper  presented  by  John  Richards  before  the  Technical  Society  of  the  Pacific 
coast  in  1904  and  published  in  the  Journal  of  the  Association  of  Engineering  Societies 
for  September,  1904,  is  the  following  account  of  the  Avery  engines,  written  by 
Professor  Sweet,  a  near  relative  of  Mr.  Avery:  "In  respect  to  the  history  of  the 
Avery  engines,  these  were  made  75  to  80  years  ago  by  William  Avery,  a  local  mechanic 
in  Syracuse.  There  were  about  50  constructed  and  put  in  use.  One  of  the  runners 
is  now  in  my  possession;  another,  that  I  saw  years  ago,  had  a  hollow  shaft  of  perhaps 
1*4 -inch  bore.  The  head  or  runner  was  of  sword  shape,  the  arm  1  by  3  inches  at  the 
center  and  %  Uy  3%  inches  at  the  ends,  the  diameter  swept  being  about  5  feet.  Steam 
was  admitted  through  the  shaft  by  means  of  a  stuffing  box,  passed  through  the  shaft  to 
the  hollow  arms  and  escaped  at  a  tangential  issue  %  inch  by  %  inch,  at  the  rear  corners 
of  each  arm,  the  ends  of  which  were  stopped  by  plugs  brazed  in.  Owing  to  the  rapid 
rotation  of  the  arms — 10  to  IS  miles  per  minute — the  front  edges  were  so  rapidly  cut 
away  that  replaceable  blades  made  of  tempered  steel  were  inserted  so  they  could  be 
renewed.  The  fact  that  the  engine  had  to  be  taken  to  a  blacksmith  shop  every  3  or  4 
months  for  renewal  or  repairs  had  more  to  do  with  its  abandonment  than  its  lack  of 
economy.  As  to  the  latter,  people  who  knew  the  facts,  or  claimed  to  do  so,  said  that 
when  they  changed  to  the  common  slide-valve  engines  there  was  no  gain  in  steam 
economy  over  the  Avery  engine.  Another  feature  that  worked  against  the  Avery 
engine  was  the  stuffing  box  around  the  shaft,  which  in  the  hands  of  workmen  of  that 
time  was  apt  to  be  set  up  so  as  to  consume  a  large  part  of  the  power  in  friction.  This 
was  a  natural  consequence,  as  the  wear  was  rapid.  What  the  result  would  have  been 
with  a  truly  ground  shaft  in  a  metal  bush,  instead  of  a  turned  shaft  and  stuffing  box, 
making  the  issues  expanding  nozzles  and  multiple  expanding  by  2  or  3  arms  in  separate 
cases  and  connecting  to  a  condenser,  is  not  known.  It  might  rival  a  pretty  good  modern 
engine,  if  not  the  best.  The  Avery  engines  were  used  in  saw  mills  and  wood-working 
shops  of  the  time.  They  had  weak  starting  power,  an*d  did  not  need  much  for  the 
uses  named.  They  ran  at  such  a  fearful  speed  that  the  reducing  motion  was  ^n  im- 
pediment. Mr.  Avery  had  to  employ  bands,  which  were  far  more  objectionable  than 
gear  wheels." 


26 


STEAM  TURBINES 


Leroy  is  perhaps  the  first  on  record  with  this  idea  of  the  applica- 
tion of  the  expanding  nozzle.  He  was  a  prolific  inventor  and  had 
definite  notions  about  many  features  now  employed  in  turbines. 
Figs.  3,  4  and  5  show  three  styles  of  rotating  arms  that  he 
proposed  for  reaction  wheels.  The  nozzle  at  N  is  clearly  a  diverg- 
ing nozzle,  as  are  also  the  orifices  in  Fig.  5.  It  is  uncertain,  how- 
ever, whether  he  understood  the  principle  of  the  diverging  nozzle, 
because  he  states  in  one  place  that  a  nozzle,  in  the  form  of  a  tube, 
Fig.  4,  will  produce  a  higher  steam  velocity  than  a  funnel-shaped 
opening.  This  would  be  true  if  the  funnel  flared  too  much,  as 


Fig.  3.  Fig.  4. 

LeRoy's  Reaction  Wheels. 


Fig.  5. 


seems  to  be  the  case.  It  is  a  curious  fact  that  the  author  has 
learned  of  recent  experiments  by  a  mechanical  engineer  who  is  at 
work  upon  the  turbine  problem,  which  show  the  same  result. 
When  a  nozzle  flares  too  much  the  expansion  of  the  steam  is  com- 
pleted before  the  end  of  the  nozzle  is  reached,  and  the  effect  of  the 
divergence  beyond  that  point  is  to  check  the  flow  of  velocity  just  as 
is  the  case  in  a  water  nozzle  which  diverges. 

Leroy  was  one  of  the  first  to  propose  a  compound  turbine.  He 
shows  two  illustrations  of  machines — one  a  reaction  and  one  an 
impulse  turbine,  in  which  each  wheel  is  encased  in  a  separate 
chamber.  In  the  reaction  turbine  steam  enters  the  hollow  arms 
of  the  first  wheel  through  a  trunnion  at  the  center  and  escapes 


EARLY  TURBINE  PATENTS 


27 


through  openings  in  the  periphery  into  the  first  chamber.  It  is 
then  conducted  by  a  pipe  to  the  second  wheel  in  a  similar  manner, 
where  it  finally  escapes  into  the  second  chamber,  and  so  on.  His 
compound  impulse  turbine  is  entirely  similar  in  principle  to  the 
Real  and  Pichon  turbine  except  that  instead  of  a  succession  of 
openings  for  the  steam  around  the  periphery  the  steam  is  con- 
ducted to  each  wheel  by  a  single  pipe.  His  drawings  of  the  com- 
pound turbine  are  unpractical,  because  he  makes  no  provision  for 
the  increasing  volume  of  the  steam  as  it  expands.  The  drawings 
show  passages  of  the  same  area  near  the  exhaust  end  of  the  tur- 
bine as  at  the  inlet  end. 


Fig.  6.  Fig.  7. 

Pilbrow's:    Wheels  rotate  in  opposite  directions. 

Another  early  inventor  who  attempted  to  use  the  diverging  noz- 
zle was  Von  Rathen,  who,  in  1847,  invented  a  reaction  wheel 
having  conical-shaped  mouthpieces  through  which  the  steam 
escaped.  These  also  flared  so  much  as  to  be  a  hindrance  rather 
than  a  help. 

Pilbrow,  1842. — The  inventions  of  Pilbrow  were  numerous.    He 


28  STEAM  TURBINES 

experimented  on  the  flow  of  steam  and  determined  that  for 
economical  results  the  peripheral  velocity  of  the  wheel  must  be 
very  high,  and  accordingly  devised  various  arrangements  for  com- 
pounding with  a  view  to  reducing  the  velocity  to  a  practical  rate. 
In  all  his  compound  turbines,  however,  he  adopted  the  plan  of  run- 
ning two  or  more  wheels  in  opposite  directions  without  stationary 
guide  vanes,  as  shown  in  Fig.  6.  Here  steam  enters  through  the 
nozzle,  A,  impinging  against  the  blades  of  wheel  C,  which  rotates 
in  the  direction  of  the  arrow.  The  steam  then  passes  through  this 
wheel  and  discharges  against  the  blades  of  a  second  wheel,  E,  ro- 
tating in  the  opposite  direction.  Fig.  8  shows  how  he  proposed 
to  carry  the  idea  still  further  by  using  several  wheels,  the  alternate 
wheels  rotating  in  opposite  directions.  Still  another  construction 


Fig.   8.     Pilbrow's   Multi- Wheel  Turbine. 

that  he  proposed  is  indicated  in  Fig.  7,  where  the  two  wheels 
rotate  on  parallel  shafts,  as  shown,  and  have  inclined  vanes  so 
located  that  steam  from  the  nozzle  will  flow  through  the  vanes  of 
both  wheels  in  the  direction  of  the  arrows.  The  buckets  are 
curved,  as  in  Fig.  6,  and  the  wheels,  of  course,  rotate  in  opposite 
directions. 

Another  interesting  invention  of  Pilbrow  is  illustrated  in  Fig.  9. 
This  is  a  reversing  turbine  arranged  with  a  number  of  nozzles  that 
can  be  shut  off  or  opened  successively  by  means  of  a  rotary  valve. 
The  plan  of  using  several  nozzles,  which  are  brought  into  or  out  of 
action  by  valves,  as  used  in  the  De  Laval  and  Curtis  turbines, 
probably  here  has  its  introduction,  and  the  invention  is  of  value  on 
this  account.  Steam  enters  the  chamber,  C,  in  which  is  located  n 


EARLY  TURBINE  PATENTS 


29 


rotating  segment  that  covers  or  uncovers  the  nozzle  openings, 
a,  b,  c,  etc.  At  A  is  the  wheel  with  vanes  pointing  in  one  direction 
and  at  B  one  with  vanes  in  the  opposite  direction.  Half  the  nozzles 
connecting  with  chamber  C  direct  the  flow  of  steam  against  wheel 
A  and  the  other  half  against  wheel  B.  By  rotating  the  segment, 
steam  can  be  admitted  to  either  wheel,  causing  the  turbine  to  re- 
volve in  either  direction,  as  desired ;  and  also  the  amount  of  steam 
admitted  can  be  adapted  to  the  power  required.  A  rotating  valve 
of  this  description  is  not  to  be  advocated  as  a  durable  construction. 


Fig.  9.     Pilbrow's  Plan  for  Reversing  with  Valve  for  Controlling  Nozzles. 

Wilson,  1848. — The  inventions  of  Wilson  rank  among  the  two 
or  three  most  important  early  steam  turbine  patents.  His  designs 
are  the  forerunners  of  the  present  Parsons  type.  He  devised  sev- 
eral compound  reaction  turbines  in  which  the  steam  flowed  through 
alternating  sets  of  stationary  and  rotating  rings  of  blades,  expand- 
ing gradually  during  its  passage  through  the  apparatus.  Fig.  10 
is  a  sketch  of  his  most  valuable  invention.  Steam  enters  at  the  left, 
passes  through  the  turbine  in  a  longitudinal  direction  and  exhausts 
at  the  outlet  at  the  right.  The  vanes,  a,  b  and  c,  are  attached  to 
the  drum,  D,  and  rotate  with  it,  while  d,  e  and  /  are  stationary 
guide  vanes.  The  depth  of  the  vanes  mcreases  from  inlet  to  outlet, 


30 


STEAM  TURBINES 


Fig.   10.     Wilson's   Compound  Turbine. 

allowing  for  gradual  expansion  of  the  steam.  This  is  really  the 
Parsons  turbine  reduced  to  its  simplest  elements. 

Another  type — the  radial  flow  wheel — is  shown  in  Fig.  11.  Here 
there  are  alternating  stationary  and  moving  vanes,  and  the  steam 
flows  outwardly  through  them,  at  the  same  time  expanding  to  a 
lower  pressure. 

In  Fig.  12  is  still  another  type  in  which  there  is  a  single  rotating 


Pig.    11.     Wilson's   Compound   Radial-flow  Turbine. 


EARLY  TURBINE  PATENTS 


31 


ring  of  blades  marked  B.  The  steam  is  expanded  and  utilized 
upon  this  one  ring  of  blades  several  times  in  succession  by  follow- 
ing a  tortuous  course  back  and  forth  through  this  ring  B.  Steam 
enters  at  A,  passes  through  the  moving  blades  to  the  chamber  C, 
then  returns  through  the  guide  vanes  in  this  chamber  to  the  cham- 
ber D;  again  it  passes  through  the  guide  vanes  to  the  wheel  and 
into  chamber  E;  then  to  chamber  F,  and  so  on.  These  successive 
chambers  increase  in  size  to  allow  for  the  increase  in  the  volume  of 


Fig.   12.     Another  Type  of  Wilson  Turbine. 

the  steam  as  it  progresses  through  the  wheel,  until  finally  it  has 
passed  around  the  whole  circumference  and  exhausts  at  the  outlet 
G.  This  plan  of  allowing  steam  to  act  at  different  points  in  suc- 
cession on  a  single  rotating  ring  of  blades,  has  since  been  worked 
out  in  various  other  ways,  as  subsequent  patent  specifications  show. 
Delonchant,  1853. — The  speed  reduction  problem  was  attacked 
by  Delonchant  in  the  same  way  that  it  was  later  by  De  Laval ;  that 
is,  instead  of  compounding  he  proposed  to  allow  his  turbine  to  run 
at  high  speed  and  then  used  reduction  gearing,  in  the  form  of  the 
familiar  "grindstone  bearing."  The  arbor,  B,  of  the  wheel,  Fig.  13, 


32 


STEAM  TURBINES 


Fig.    13.     Delonchant. 

was  supported  on  the  circumference  of  antifriction  wheels,  C.  In 
explanation  he  says :  "By  the  employment  of  these  wheels  instead 
of  ordinary  bearings,  not  only  the  rubbing  of  the  first  axes  will  be 
replaced  by  rolling  friction  but  power  will  also  be  transmitted  to 
the  following  parts,  without  gearing."  In  the  illustration,  A  is  the 
rotating  wheel ;  B,  the  arbor  and  at  the  center  is  a  steam  chest,  D  D, 
indicated  in  outline  only.  Steam  passes  from  the  steam  chest 
through  the  passages  d  d  d;  and  £  is  a  ring  having  passages  c  e  e, 
used  in  regulating  the  amount  of  steam  flowing  through  the  wheel. 
The  passages,  d  d  d,  are  so  disposed  that  by  rotating  ring  E  the 
passage  e  e  e  through  the  ring  will  be  successively  cut  off  from  the 
steam  supply,  or  else  opened  to  the  supply.  By  moving  ^12  °f  a 
turn  one  passage  is  closed;  another  %2  closes  a  second  passage, 
and  so  on. 

Tournaire,  1853. — In  this.  vear  Tournaire  presented  to  the 
Academic  des  Sciences  a  paper  discussing  the  merits  of  compound 
turbines  both  of  the  impulse  and  reaction  types.  There  is  a  copious 
extract  from  this  paper  in  the  Bulletin  de  la  Societe  d'Encourage- 
ment  pour  Tlndustrie  Nationale  for  September,  1896,  and  the  facts 
explained  by  him  as  essential  to  a  successful  turbine  are  so  in 
accordance  with  modern  practice  as  to  place  him  among  the  leading 


EARLY  TURBINE  PATENTS  33- 

inventors.  He  says :  "To  overcome  the  difficulties  of  high  veloci- 
ties the  vapor  or  gas  should  be  made  to  lose  its  pressure  in  a  con- 
tinuous and  gradual  manner,  or  by  successive  fractions,  by  causing 
it  to  react  several  times  upon  the  floats  of  turbines  conveniently 
situated.  Since  the  differences  of  pressure  are  considerable  it  is 
not  difficult  to  recognize  the  necessity  for  a  large  number  of  suc- 
cessive turbines  in  order  to  sufficiently  annul  the  velocity  of  the 
fluid  jet.  In  spite  of  the  multiplicity  of  parts  the  device  must  be 
simple  in  its  action  and  susceptible  of  great  exactness  in  construc- 
tion." Tournaire  believed  he  fulfilled  these  conditions  by  means 
of  a  machine  composed  of  several  wheels,  with  shafts  having  the 
same  axis  and  driving  the  wheel  which  was  to  transmit  the  motion, 


\\V\\v 


/G 


Fig.    14.     Plan    of    Vanes 

in  Tournaire's 

Turbine. 

by  means  of  pinions.  A  plan  of  the  buckets  and  vanes  is  given 
in  Fig.  14,  where  G  G  G  are  the  rotating  elements  and  V  V  V  the 
stationary  elements.  He  describes  the  construction  of  the  turbine 
in  detail,  but  these  structural  features  are  of  little  interest  at  the 
present  time.  It  is  to  be  noted,  however,  that  he  appreciated  fully 
the  necessity  for  expansion.  He  says  :  "As  the  vapor  will  expand 
in  proportion  as  it  passes  from  the  wheel  buckets  and  directing 
rings,  it  is  necessary  that  the  passages  between  them  become  larger 
and  larger."  He  also  suggests  losses  from  leakage,  saying:  "A 
part  of  the  fluid  escaping  between  the  spaces,  which  it  is  necessary 
to  leave  between  the  fixed  and  movable  parts,  will  exert  no  action 
upon  the  turbine,  nor  will  it  be  guided  by  the  directing  buckets. 
Shocks  and  eddies  will  be  produced  at  the  entrance  and  exits  of 


34 


STEAM  TURBINES 


the  buckets."  Again,  "The  friction  which  the  narrowness  of  the 
channel  will  render  considerable  will  absorb  an  appreciable  part  of 
the  theoretical  work."  As  to  the  structural  features  he  suggests 
among  other  things,  that  "the  cogs  of  the  pinions  which  will  turn 
with  great  rapidity,  should  work  very  evenly  without  shocks  and 
jolts,"  and  proposes  the  use  of  helicoidal  gears.  His  turbine,  as 

well  as  some  of  the  others  already  de- 
scribed, is  a  vertical  turbine  rotating 
on  a  vertical  axis. 

John  and  Ezra  Hartman,  1858. — • 
Standing  in  importance  with  the  inven- 
tions of  Wilson  and  Tournaire  are  the 
English  and  American  patents  of  the 
Hartman  brothers,  from  the  drawings 
of  which  Fig.  15  is  made.  The  patent 
relates  to  a  "mode  of  obtaining  motive 
power  by  causing  steam  or  air  to  im- 
pinge upon  a  series  of  chambers  with 
curved  bottoms  ranged  around  a  wheel 
at  or  near  the  periphery  thereof;  and 
second,  the  general  construction  and 
arrangement  of  machinery  or  ap- 
paratus for  obtaining  motive  power." 
Fig.  15  shows  the  most  important 
modification  of  the  patent,  of  which 
the  following  is  the  inventor's  descrip- 
tion :  "This  represents  a  detail  of  the 
third  modification  wherein  we  propose 
to  employ  two  wheels,  C  C1,  both 
wheels  being  fast  on  one  shaft,  D. 

A  space  is  left  between  the  contiguous  faces  of  these  wheels  for  the 
reception  of  four  or  more  returning  chambers,  d  d,  the  bottoms  of 
which  are  curved  in  a  direction  opposite  to  that  of  the  bottoms  of 
the  chambers  in  the  wheels.  These  chambers  in  other  respects  are 
precisely  similar  to  those  in  the  wheels  and  are  fitted  to  a  rim  which 
is  bolted  or  otherwise  secured  to  the  interior  of  the  casing  G.  The 
jet  pipe,  F,  is  at  one  side  of  the  wheel  and  the  discharge  pipe,  H,  on 
the  opposite  side  of  the  second  wheel. 


Fig. 


15.     Hartman's    Compound 
Impulse  Turbine. 


EARLY  TURBINE  PATENTS 


35 


"The  jet  pipe  on  being  first  introduced  impinges  against  the 
curved  bottoms  of  the  chambers  in  the  wheel  C,  and  is  thence  di- 
verted against  the  fixed  chambers,  d  d,  whence  it  is  again  diverted 
on  to  the  curved  bottoms  of  the  chambers  in  the  second  wheel,  C, 
and  finally  passes  off  by  the  escape  pipe,  H" 

Charles  M orison,  1862. — We  have  already  illustrated  types  of 


ROTATING 
ARMS 


Fig.  16.     Monson's  Compound  Reaction  Wheel. 

simple  reaction  wheels,  but  a  search  of  the  patent  records  shows 
that  several  inventors  have  attempted  to  improve  on  this  arrange- 
ment and  produce  a  turbine  which  will  run  at  slower  speed,  by 
having  a  succession  of  simple  reaction  wheels,  each  one  in  a  sepa- 
rate chamber  and  arranged  so  that  steam  issuing  from  a  wheel  into 
its  chamber  will  then  pass  through  to  the  next  wheel,  and  so  on. 
This  in  substance  is  the  design  of  Monson's  turbine,  shown  in  Fig. 
16.  The  leading  specification  of  his  patent  is  as  follows :  "A  re- 
peating rotary  engine  constructed  in  a  manner  so  as  to  operate  sub- 
stantially as  described  ;  namely,  of  two  or  more  sets  of  curved  arms, 
B,  C,  D,  or  their  mechanical  equivalents ;  a  series  of  two  or  more 
tight  chambers  or  passages,  A,  and  a  shaft  or  its  equivalent  divided 


36 


STEAM  TURBINES 


into  separate  chambers  and  provided  with  induction  and  escape  pas- 
sages." The  course  followed  by  the  steam  will  be  evident  from  the 
engraving.  Other  patents  similar  to  Monson's  have  been  taken 
out  in  later  years,  notably  by  T.  Bauta  in  1867  and  by  Parsons,  the 
inventor  of  the  Parsons  steam  turbine,  in  1893.  These  latter  are 
exactly  similar  in  principle  to  Monson's  and  are  merely  construc- 
tion patents. 


Fig.  17.     Hoehl,  Brakell  &  Gunther. 

Hoehl,  Brakell  &  Gunther,  1863. — The  turbine  produced  by  this 
aggregation  of  inventors  has  as  its  only  novelty  an  arrangement  of 
passages  by  which  the  steam  returns  on  itself  and  so  is  utilized 
twice  by  the  same  wheel,  although  an  additional  set  of  wheel 
blades  is  required.  Steam  enters  the  chamber  A,  passes  radially 
outward  through  the  guide  passages  B  to  the  wheel  blades  C,  then 
discharges  into  the  annular  chamber  D,  where  its  motion  is  re- 
versed, and  escapes  through  E  to  another  set  of  buckets  F.  and 
finally  to  the  exhaust  chamber  G. 

Perrigault  &  Farcot. — The  type  of  turbine  here  exemplified  is  in 
principle  like  Wilson's  turbine,  in  Fig.  12,  the  latest  representative 
of  which  is  found  in  the  compound  turbine  of  Messrs.  Riedler  and 
Stumpf,  to  which  reference  will  be  made  later.  The  inventions  of 
Perrigault  &  Farcot  took  several  forms,  but  the  general  principle  is 
well  illustrated  by  Fig.  18  herewith.  Here  steam  enters  through 
the  pipe  or  nozzle  A  and  impinges  against  the  wheel  buckets, 
passing  through  to  the  other  side  of  the  wheel  where  it  discharges 


EARLY  TURBINE  PATENTS 


37 


into  pipe  B.  This  pipe  brings  the  steam  around  again  to  the  inlet 
side  of  the  wheel,  allowing  it  to  discharge  a  second  time  against  the 
buckets  of  the  same  wheel,  when  it  is  again  picked  up  by  a  second 
pipe  C,  and  so  on.  The  exhaust  is  finally  through  pipe  D. 

The  arrangement  consists  essentially  in  a  bundle  of  bent  pipes 
having  openings  a,  b,  c,  through  which  the  steam  impinges  against 
the  wheel  buckets;  and  openings  on  the  other  side,  x,  y,  etc., 
which  gather  up  the  steam  flowing  from  the  wheel  and  bring  it 
round  again  to  the  inlet  side.  The  object  is  to  utilize  the  steam 


a   v     c 
Fig.    18.     Compound  Turbine   with   Only   One  Wheel. 

over  and  over  without  necessitating  a  series  of  rotating  wheels. 
It  is  a  system  that  has  been  tried  with  various  modifications,  but 
without  much  success.  Its  obvious  disadvantages  are:  Large 
losses  from  friction  and  leakage,  and  the  wide  range  of  tempera- 
tures through  which  the  same  buckets  must  pass,  thus  causing 
condensation  and  reevaporation  as  in  the  steam  engine. 

Moorhouse,  1877. — Among  the  most  successful  types  of  turbine 
is  that  having  a  succession  of  chambers  in  each  of  which  is  a  single 
impulse  wheel.  There  is  only  a  slight  drop  in  pressure  from  cham- 
ber to  chamber,  so  that  the  velocity  of  the  steam  does  not  become 
excessively  high  at  any  point.  The  latest  turbine  of  this  description 
is  the  Hamilton-Holzwarth,  built  by  the  Hooveris,  Owens,  Rent- 
schler  Co.,  Hamilton,  O.,  and  the  earliest  one  of  which  there  is 
any  record  is  the  invention  of  Real  &  Pichon,  1827 — the  first  pa- 
tent referred  to  in  this  series.  The  invention  of  Moorhouse  is  for 


38 


STEAM  TURBINES 


a  turbine  on  the  same  plan.  As  shown  in  Fig.  19,  a,  a,  a,  etc.,  are 
the  nozzles,  and  b,  b,  b,  etc.,  the  wheel  buckets.  Steam  flows 
radially  outward  at  each  wheel  until  near  the  exhaust  end,  where 
there  is  a  different  arrangement,  owing  to  a  smaller  drop  in  pres- 
sure between  the  successive  chambers. 

Moorhouse  realized  what  previous  inventors  of  this  type  of 
compound  impulse  turbine  had  not,  or  at  least  had  failed  to  specify, 
namely,  that  provision  must  be  made  for  progressive  expansion  of 
the  steam  by  a  gradual  increase  in  the  area  of  the  steam  passages. 


a  a  a 


Fig.  19.     Compound  Impulse  Turbine  with  Provision  for  Expansion  of 
the  Steam. 

His  patent  is  a  broad  one  and  fully  covers  this  requirement  as  a 
general  principle,  apart  from  the  exact  method  used  in  its  applica- 
tion. In  his  specifications  he  says  that  his  invention  consists  of  a 
cylinder  built  in  sections,  each  section  composing  a  separate  com- 
partment. Through  the  center  of  the  cylinder  passes  a  revolving 
driving  shaft  upon  which  is  fitted  a  series  of  turbine  wheels,  each 
having,  at  or  near  its  circumference,  a  sufficient  number  of  buckets. 
These  wheels  are  as  many  in  number  as  the  compartments  into 
which  the  cylinder  is  divided.  Each  compartment  contains  a  tur- 
bine wheel,  and  is  separated  from  the  adjoining  compartment  by 
means  of  a  dividing  plate  or  diaphragm.  The  foregoing  is  con- 
densed from  his  specifications,  but  correctly  represents  their  mean- 
ing. He  then  goes  on  to  say : 


EARLY  TURBINE  PATENTS 


39 


"Openings  are  made  in  the  dividing  plates  which  separate  each 
compartment  from  the  adjoining  ones,  and  the  area  of  these  open- 
ings is  proportioned  to  the  pressure  of  the  steam  or  other  driving 
fluid,  and  to  the  number  of  compartments  and  turbine  wheels,  and 
to  the  extent  to  which  it  is  desired  that  the  driving  fluid  should  be 
expanded  before  being  finally  discharged  from  the  engine.  By  this 
means  the  driving  fluid,  admitted  at  its  highest  pressure  into  the 
smallest  compartment,  passes  into  the  second  compartment  through 


Fig.    20.     Radial    Outward-flow    Turbine. 

openings  of  such  area  that  it  expands  to  a  calculated  extent.  The 
same  process  is  repeated,  etc." 

He  says  that  various  forms  of  turbine  may  be  used,  and  his  first 
claim  is  as  follows : 

"In  combination  with  a  rotary  engine,  the  dividing  plates  be- 
tween the  compartments  provided  with  openings  forming  com- 
munications respectively  of  varying  area  between  said  compart- 
ments, the  turbine  wheels  in  such  compartments,  and  a  driving 
shaft,  substantially  as  and  for  the  purpose  set  forth." 

Cutler,  18/9. — This  is  a  radial  outward  flow  turbine  in  which 
the  compound  principle  is  used.  Steam  enters  at  the  bottom,  passes 
to  the  center  and  then  flows  radially  outward  through  the  passages 


40 


STEAM  TURBINES 


between  the  guides  c  c  c  and  the  wheel  vanes  d  d  d.  The  rotating 
wheel  A  has  vanes  attached  to  each  of  its  two  faces  so  the  pressure 
is  balanced  on  each  side.  Expansion  of  the  steam  is  allowed  for  in 
part  by  the  increasing  width  of  the  passages  and  in  part  by  the 
fact  that  the  steam  is  constantly  flowing  from  a  smaller  to  a  larger 
diameter  of  wheel  so  that  the  circumferential  area  of  the  passages 
constantly  increases  from  this  cause  as  well. 

Imray,  1881. — This  is  another  attempt  to  apply  the  compound 


SECTION  ON  X-Y      v 

Fig.  21.     Another  Compound  Turbine  with   Only  One   Wheel. 

principle  to  a  single  wheel,  having  a  single  set  of  vanes,  but  differs 
somewhat  from  the  turbines  of  Wilson  and  Perrigault  and  Farcot. 
Steam  enters  at  A,  passes  through  the  nozzle  and  impinges  against 
the  buckets  C  C  C.  These  buckets  are  semi-circular  in  shape,  as 
indicated  in  the  sectional  view  in  the  upper  left-hand  corner  of  the 
illustration.  The  steam  enters  at  one  side  of  the  bucket,  follows 
the  curved  surface  of  the  bucket,  and  discharges  into  the  opposite 
side  of  a  semi-circular  stationary  bucket  or  guide  D.  Here  the 
direction  of  flow  of  the  steam  is  again  reversed.  The  steam,  as 
before,  flows  around  the  stationary  guide  surface  and  discharges 
against  one  of  the  buckets  C,  whence  it  is  carried  along  to  the 
second  stationary  bucket,  and  so  on,  alternately  entering  the  succes- 


EARLY  TURBINE  PATENTS 


41 


sive  wheel  buckets  C  C  C  and  the  successive  stationary  buckets 
D  D  D.  It  finally  discharges  on  the  opposite  side  of  the  turbine 
casing,  at  E.  In  the  meantime  steam  enters  at  A  on  the  right-hand 
side  of  the  casing  and  zigzags  through  the  lower  half  of  the  wheel 
in  a  similar  manner,  exhausting  at  E  on  the  left-hand  side. 

De  Laval,  1883. — The  first  patent  of  this  noted  inventor  was  for 
a  reaction  turbine  and  was  taken  out  in  1883,  in  several  countries. 
According  to  the  specifications,  steam  (or  other  fluid)  enters  the 


Fig.  22.     Two  Wheels   Rotating  in  Opposite  Directions. 

wheel  at  the  center  through  a  nozzle,  and  passes  outward  through 
hollow  curved  arms,  escaping  at  their  ends,  and  causing  the  wheel 
to  rotate  at  high  velocity.  The  wheel  shaft  drives  another  shaft 
at  a  slower  speed,  by  means  of  friction  wheels,  the  requisite 
pressure  between  the  friction  surfaces  being  obtained  by  the  axial 
thrust  of  the  turbine  wheel.  The  principle  of  this  turbine  is  no  dif- 
ferent from  that  of  the  first  American  patent  by  Avery  in  1831,  but 
its  application  to  centrifugal  cream  separators,  for  the  extensive 
'development  of  which  Dr.  De  Laval  has  been  responsible,  was 
successful  and  marks  the  beginning  of  an  important  career  by  this 
inventor  in  the  manufacture  of  steam  turbines. 

Babbitt,   1884. — B.   T.   Babbitt,   besides   acquiring  fame   as   a 


42 


STEAM  TURBINES 


manufacturer  of  laundry  soap,  was  both  an  inventor  and  a  me- 
chanic, and  one  of  his  inventions  related  to  a  steam  turbine  of  the 
type  shown  in  Fig.  22.  This  is  an  inward-flow  turbine  with  two 
wheels,  A  and  B,  having  rows  of  buckets  on  their  peripheries.  The 
wheels  rotate  in  opposite  directions.  The  steam  from  the  nozzle  N 
impinges  against  the  buckets  of  the  outer  wheel,  which  it  is  sup- 
posed to  leave  with  a  considerable  residual  velocity,  and  gives  up 
its  remaining  energy  to  the  buckets  of  the  inner  wheel.  The  chief 
novelty  of  the  invention  is  the  method  of  transmitting  power  from 
these  two  wheels  to  the  slow  speed  shaft  F.  The  turbine  wheels 

are  mounted  concentrically  on  two  sepa- 
rate shafts,  having  the  same  axis,  shown 
at  E.  There  is  a  pinion  on  the  outer  end 
of  each  of  the  shafts.  One  of  these 
pinions  gears  with  the  internal  gear  C  on 
one  end  of  the  slow  speed  shaft  F,  and 
the  other  pinion  gears  with  the  spur  wheel 
D  on  the  other  end  of  shaft  F. 

Isaac  Last,  1885. — This  is  another  ex- 
ample where  the  steam  is  caused  to  re- 
turn upon  itself,  first  flowing  radially  out- 
ward then  reversing  and  flowing  radially 
inward,  just  as  in  the  Hoehl,  Brakell  & 
Gunther  turbine,  Fig.  17.  In  Last's  tur- 
bine, however,  compounding  is  carried 
further  than  in  the  former,  there  being  a 

series  of  wheels  placed  side  by  side  on  the  same  shaft.  In  the  sketch, 
ABC  are  the  different  chambers  in  which  the  wheels  rotate ;  a  a 
are  the  guides  for  directing  the  steam  against  the  wheel  buckets 
and  b  b  are  the  wheel  buckets.  One  drawing  in  the  patent  speci- 
fications of  Last  has  a  very  modern  appearance,  in  that  he  shows 
a  compound  turbine  built  up  of  two  parts,  the  high  pressure  and 
the  low  pressure,  in  each  of  which  is  a  series  of  compound  wheels. 
The  high-pressure  and  low-pressure  sections  are  connected  by  a 
pipe,  and  their  arrangement  resembles  that  of  some  of  the  turbines 
built  to-day. 

Parsons,  1885. — With  a  patent  issued  in  several  countries  in  this 
year,  the  Hon.  C.  A.  Parsons,  who  was  the  first  to  place  the  turbine 


Fig.    23.     Isaac    Last. 


EARLY  TURBINE  PATENTS 


43 


on  a  commercial  basis,  enters  the  field.  In  all  his  work  he  has 
adhered  to  the  reaction  turbine  and  is  responsible  for  the  success- 
ful development  of  the  compound  reaction  motor.  In  Fig.  24  is 
a  section  of  one  of  his  patent  drawings.  Steam  enters  at  the  center 
A,  and  passes  right  and  left  between  the  series  of  guide  vanes  at- 


Fig.    24.     Parson's    First    Important    Patent. 


Fig.   25.     Bearing  for   Parson's   Turbine. 

tached  to  the  outer  casing  and  the  rotating  blades  attached  to  the 
inner  drum,  which  has  the  journals,  B  B.  Steam  escapes  to  the 
exhaust  passages  E  E  and  finally  to  the  exhaust  pipe  shown  by 
dotted  lines.  The  following  paragraphs  are  extracts  from  his  de- 
scription : 

"I  arrange  the  portions  of  the  motor  to  form  an  approximately 
cylindrical  figure,  the  whole  being  mounted  upon  the  same  shaft, 


44  STEAM  TURBINES 

the  first  delivering  into  the  second,  the  second  into  the  third  and  so 
on.  Each  portion  comprises  a  set  of  fixed  and  a  set  of  moving 
vanes,  the  direction  of  motion  of  the  actuating  fluid  being  generally 
parallel,  or  approximately  so.  To  balance  the  end  pressure  upon 
the  cylinder,  I  mount  two  similar  sets  of  rotary  parts  upon  one 
shaft,  one  set  being  so  placed  at  each  side  of  the  inlet  for  the  actuat- 
ing fluid  that  the  entering  stream  divides  right  and  left,  and  the  ex- 
haust takes  place  at  both  ends. 

"As  the  speed  of  the  motor  will  be  necessarily  high,  and  perfect 
balancing  of  the  moving  parts  would  not  be  practicable,  I  give  to 
the  bearings  a  certain  very  small  amount  of  elasticity  or  play  com- 
bined with  a  frictional  resistance  to  their  motion." 

This  refers  to  the  well-known  Parsons  construction  shown  in 
Fig.  25,  in  which  there  is  an  annular  space,  between  the  shell  of 
the  bearing  and  the  pocket  for  the  shell  bored  out  in  the  frame, 
filled  with  a  series  of  metal  rings.  Every  other  ring  is  bored  to  fit 
the  outside  of  the  shell,  but  its  outside  diameter  is  smaller  than  the 
bore  of  the  pocket,  as  in  a  a.  The  alternate  rings  b  b  are  turned  to 
fit  the  pocket,  but  are  bored  larger  than  the  outside  of  the  shell. 
The  rings  are  forced  together  by  a  spring  s,  so  that  they  offer  con- 
siderable resistance  to  any  lateral  movement  of  the  bearing. 

He  says :  "The  lubrication  is  effected  by  forcing  lubricant 
through  pipes  to  the  parts  to  be  lubricated  and  for  this  purpose  a 
pump  can  be  employed.  To  prevent  leakage  past  the  shaft  at  the 
end  covers  of  the  casing,  which,  when  steam  is  the  actuating  fluid, 
would  be  inconvenient,  I  form  annular  recesses  in  the  covers 
around  the  shaft  ends,  and  place  these  recesses  in  communication 
with  a  pipe  in  which  a  partial  vacuum  is  maintained  by  suitable 
means,  such  as  a  steam  jet." 

In  the  next  year,  1888,  Parsons  took  out  a  patent  in  which  the 
turbine  wheels  are  arranged  in  groups,  each  successive  group  being 
of  larger  diameter  than  the  preceding  one,  to  allow  the  steam,  as 
it  expands,  to  flow  through  larger  spaces,  as  required  by  the  in- 
crease in  the  specific  volume  of  the  steam.  He  also  proposes  to 
secure  steam-tight  joints  at  the  bearings  by  admitting  water  under 
pressure  to  an  annular  groove  passing  around  the  bearing.  He 
suggests  cutting  a  spiral  groove  on  the  shaft,  at  the  section  where 
this  annular  groove  occurs,  with  the  idea  that  when  the  shaft  is 


EARLY  TURBINE  PATENTS 


45 


revolving  at  high  speed,  the  spiral  will  diminish  the  quantity  of 
water  forced  into  the  turbine  casing  by  the  air  pressure,  when  the 
turbine  is  running  condensing. 

Altham,  1892. — A  compound  turbine  consisting  of  two  rotating 
wheels,  one  inside  of  and  concentric  with  the  other,  is  the  invention 
of  George  J.  Altham.  The  buckets  of  the  inner  wheel  are  arranged 
in  its  outer  periphery  and  those  of  the  outer  wheel  in  its  inner 
periphery,  so  that  steam  will  act  alternately  on  the  inner  and  outer 


Fig.    26.     Altham    Two-wheel   Turbine. 

wheel,  and  successively  on  the  different  buckets  of  both  wheels, 
the  arrangement  being  such  that  the  wheels  rotate  simultaneously 
in  opposite  directions.  Fig.  26  shows  at  the  left  a  cross-section 
of  the  rims  of  the  two  wheels  in  which  the  buckets  are  cut.  The 
other  sketches  show  longitudinal  sections  of  the  rims.  Steam  is 
discharged  from  the  nozzle  into  one  row  of  buckets  of  the  outer 
wheel,  whence  it  passes  to  the  first  row  of  buckets  of  the  inner 
wheel,  thence  to  the  second  row  of  the  outer  wheel  and  finally  to 
the  last  row  of  the  inner  wheel,  from  which  it  discharges  into  the 
turbine  casing.  In  the  smallest  sketch,  Fig.  26,  the  construction 
is  indicated  where  there  is  only  a  single  row  of  buckets  in  each 
wheel.  In  this  same  year  patents  were  granted  to  J.  F.  McElroy 


46 


STEAM  TURBINES 


for  a  turbine  with  U-shaped  channels,  but  with  one  set  of  vanes 
attached  to  the  casing.    See,  also,  Imray's  patent  of  1881. 

Dow,  1893. — When  turbines  first  began  to  come  into  promi- 
nence in  this  country,  the  one  invented  by  J.  H.  Dow  was  one  of 
the  three  or  four  that  were  most  frequently  mentioned.  His  first 
patent  was  issued  in  1887,  and  later  several  others  were  taken  out, 
but  the  one  showing  the  most  completely  worked  out  design  was 
issued  in  1893.  All  of  the  Dow  turbines  are  of  the  radial  outward- 
flow  type,  consisting  of  alternating  rings  of  rotating  and  stationary 
vanes,  and  in  this  respect  resemble  one  of  Wilson's  inventions  of 


Fig.  27.     Dow's  Patent  of  1893. 

1848.  In  Fig.  27,  taken  from  his  latest  patent,  A  is  the  ring  of 
stationary  vanes  directing  the  steam  against  the  ring  of  rotating 
blades  B,  and  C  is  another  ring  of  stationary  vanes,  D  a  ring  of 
rotating  blades,  etc.  A  peculiarity  of  the  drawing  shown  is  that 
the  stationary  vanes  are  not  curved  at  their  inlet  ends  in  a  way  to 
guide  the  steam  into  them  in  the  direction  in  which  it  leaves  the  ro- 
tating blades,  except  as  the  latter  might  be  designed  so  that  steam 
would  leave  them  in  the  direction  in  which  the  wheel  is  turning, 
which  would  be  an  inefficient  arrangement.  As  actually  con- 
structed, however,  the  guide  vanes  were  curved  correctly  and  the 
turbine  was  built  along  the  lines  shown  in  the  reference  to  it  in 
Thurston's  manual  of  the  steam  engine.  In  the  patent  of  1887 
there  is  a  single  shaft  on  which  are  two  disks  facing  each  other, 
having  annular  rows  of  vanes  cut  on  their  inner  faces.  Between 


EARLY  TURBINE  PATENTS 


47 


these  two  disks  is  a  central  stationary  disk  with  annular  rows  of 
guide  vanes  cut  on  each  of  its  faces.  The  arrangement  is  shown 
in  Fig.  28.  Steam  enters  at  the  center,  and  flows  radially  outward 
between  the  vanes  on  each  side  of  the  central  disk.  In  his  latest 
patent  Dow  compounds  his  turbine  still  further  by  providing  sev- 
eral rotating  and  stationary  disks  ranged  along  the  shaft  on  each 
side  of  the  center.  Steam  enters  at  the  center  and  gradually  works 
outward  toward  both  ends  of  the  turbine. 


Fig.    28.     Dow  Turbine    with   Two 
Disks. 


Fig.     29.     De     Laval    Turbine 
with  Diverging  Nozzle. 


De  Laval,  1894. — De  Laval's  most  important  patent  relates  to  his 
expanding  nozzle,  in  combination  with  a  turbine  wheel.  It  is  in- 
teresting to  note  in  this  connection  that  the  expanding  nozzle  was 
patented  in  this  country  in  1867,  patent  64,539,  for  steam  injectors. 
De  Laval,  however,  was  the  first  to  apply  the  principle  of  a  diverg- 
ing nozzle  for  the  expansion  of  steam  to  a  turbine.  The  two 
broadest  claims  of  the  patent  are  the  following : 

1.  The  combination  with  a  bucket  or  turbine  wheel,  of  a  sta- 
tionary nozzle  opening  adjacent  to  the  wheel  and  having  its  bore 
diverging  or  increasing  in  area  of  cross  section  toward  its  dis- 
charge end,  whereby  the  elastic  fluid  under  pressure  is  expanded 
in  passing  through  the  diverging  nozzle  and  its  pressure  is  con- 
verted into  velocity  before  the  jet  is  delivered  against  the  wheel. 


48 


STEAM  TURBINES 


2.  The  combination  with  a  bucket  or  turbine  wheel,  of  a  sta- 
tionary nozzle  opening  adjacent  to  the  wheel  and  provided  with  a 
contracted  receiving  portion  and  with  a  discharge  portion  having 
its  bore  diverging  or  increasing  in  area  or  cross  section  toward  its 
discharge  end. 

Maison  Breguet,  1894. — Judging  from  the  illustration  accom- 
panying this  patent,  it  introduces  no  new  principle  that  was  not 
included  in  the  invention  of  Hartman's  compound  impulse  turbine, 
the  patent  for  which  was  taken  out  in  1858.  That  is  to  say,  the 


Fig.    30.     From    Patent   issued   to   the   firm 
of   Breguet,   Paris. 

illustration  shows  a  converging  nozzle  in  connection  with  rotating 
rings  of  blades  alternating  with  stationary  vanes.  This  is  nothing 
more  or  less  than  what  is  shown  in  Hartman's  patent  drawings. 
From  the  way  the  text  of  the  Breguet  patent  reads,  however,  the 
inventor  apparently  had  in  mind  the  improvement  of  the  De  Laval 
turbine,  and  if  such  is  the  case  he  evidently  intended  to  imply  the 
use  of  a  diverging  nozzle  instead  of  a  converging  nozzle  in  con- 
nection with  a  compound  turbine.  Putting  this  interpretation 
upon  the  patent  it  is  of  importance  as  the  first  to  be  issued  upon  this 
combination  of  elements,  preceding,  as  it  does,  the  Curtis  patent 
(which  introduced  the  same  principle)  by  about  two  years.  The 
description  of  the  invention  states  that  in  the  De  Laval  turbine 
"even  with  a  circumferential  velocity  of  the  turbine  of  420  meters, 


EARLY  TURBINE  PATENTS 


49 


if  the  steam  has  a  velocity  of  1,100  meters,  it  still  discharges  from 
this  turbine  with  a  velocity  of  4-10  meters,  and  this  velocity  is  much 
higher  when  the  circumferential  velocity  of  the  turbine  is  less. 
The  idea  that  has  naturally  come  to  us  is  to  utilize  anew  this  lost 
velocity  in  a  second  turbine  mounted  on  the  same  axis,  and  even 
in  exceptional  cases  in  a  third,  so  as  to  increase  the  use  of  the  tur- 
bine. We  affirm  as  our  property  the  invention  of  the  compound 
steam  or  gas  turbine,  in  which  the  steam,  or  gas,  after  having  lost 
a  part  of  its  live  force  in  the  turbine  buckets,  finally  loses  the  re- 
mainder in  the  buckets  of  one  or  of  several  other  disks  mounted 
on  the  same  arbor." 


Fig.    31.     Seger's    First    Patent. 

Segcr,  1894. — Seger's  turbine  has  been  built  and  used  to  some 
extent  abroad.  His  first  patent  specification,  issued  in  1893,  shows 
an  arrangement  of  wheels  indicated  in  Fig.  31.  Like  Pilbrow  and 
B.  T.  Babbitt,  he  seeks  to  secure  a  moderate  speed  of  rotation  by 
using  two  wheels  turning  in  opposite  directions.  The  steam,  in 
leaving  the  first  wheel,  impinges  directly  against  the  second  with- 
out any  intervening  guide  vanes.  The  turbine  wheels  are  on 
separate,  parallel  shafts,  and  at  G  are  the  gears  by  which  the 
motion  of  the  wheels  is  transmitted  to  the  driving  shaft,  5.  N  is 
the  nozzle  through  which  steam  enters,  and  E  the  exhaust  pas- 
sage. His  claim  is  for  "a  steam  turbine  in  which  the  turbine 
wheels  are  placed  in  close  proximity  to  each  other,  and  are  com- 


50 


STEAM  TURBINES 


bined  with  one  or  more  steam  conduits  discharging  into  the  sides 
of  said  wheels  in  such  a  manner  that  the  steam  passes  through  the 
wheels  in  the  direction  of  their  axes,  and  in  which  the  shafts  are 
arranged  out  of  line  with  each  other  so  that  the  wheels  only  partly 
overlap  each  other." 

In  his  patent  of  1894,  Seger  shows  wheels  arranged  on  the  same 
axis,  but  rotating  in  opposite  directions.  A  feature  of  the  patent 
is  the  method  for  fastening  the  buckets  in  diagonal  slots  cut  in  the 


Fig.   32.     Arrangement  of  Wheels. 


Fig.  33.     Vanes  of  Seger  Turbine. 

wheel  rims.  The  lower  view,  Fig.  32,  shows  these  slots,  and  in  the 
upper  view  the  rings,  R,  forced  inside  the  rim,  hold  the  projecting 
ends  of  the  buckets  in  position.  This  construction  will  be  evident 
from  Fig.  33,  where  A  is  one  of  the  buckets.  At  B  the  bucket  is 
placed  in  the  rim  and  at  CC  its  projecting  ends  are  bent  over  un- 
derneath the  rim.  In  1897  Seger  issued  an  English  patent  upon  an 
arrangement  of  his  turbine  by  which  the  belted  connection  could 
be  used  for  driving  the  low-speed  shaft,  Fig.  34,  from  which  power 
is  taken.  Here  A  and  B  are  the  turbine  wheels  rotating  in  oppo- 
site directions  and  attached  to  the  ends  of  shafts  which  carry,  at 
their  outer  ends,  small  belt  pulleys.  On  the  shaft,  S,  are  two  pul- 
leys, Wlt  W2,  of  equal  diameter,  one  of  which  is  fast  to  the  shaft, 


EARLY  TURBINE  PATENTS 


51 


while  the  other  is  loose  on  the  same  shaft.  The  belt  passing  around 
these  several  pulleys,  as  indicated,  transmits  power  from  the  tur- 
bine wheels  to  the  main  shaft. 


Fig.  34.     Front  and  Side  Elevation  of  Seger  Turbine. 

McElroy,  1894. — This  year  marks  the  beginning  of  the  im- 
portant patents  on  turbine  wheels  of  the  Pelton  type  for  use  with 
an  elastic  fluid.  Two  patents  were  issued,  one  to  J.  F.  McElroy 
in  this  country,  and  one  in  England  to  Professor  A.  Rateau.  The 
arrangement  of  the  nozzles  and  wheel  of  McElroy's  invention  is 
shown  at  A  in  the  sectional  view,  Fig.  35.  Nozzles  with  diverging 
mouthpieces  are  used.  At  B  is  a  section  of  a  bucket  in  a  plane 
parallel  with  the  shaft,  together  with  an  enlarged  view  of  one  of  the 
nozzles.  At  C  is  a  section  through  a  bucket  taken  in  the  other 
direction.  While  an  efficient  type  of  bucket  is  shown,  and  Mc- 
Elroy's claims  as  to  its  shape  are  broad,  taken  by  themselves,  they 
are  combined  with  certain  other  constructive  features  which  limit 
the  scope  of  the  patent.  What  he  claims  is  first  a  wheel  comprising 
a  solid  metal  ring  having  a  series  of  inclined  pockets  therein,  the 


52 


STEAM  TURBINES 


Fig.    35.     An    Early   "Pelton"   Type. 

pockets  of  each  pair  being  divided  by  a  tapering  ridge  in  combina- 
tion with  a  circular  steam  ring  having  a  circular  series  of  nozzles ; 
and  secondly,  pockets  as  above  described,  but  with  a  flat  inclined 
cut-away  portion  for  each,  as  shown  in  the  sketch. 

Rateau,  1894. — Professor  Rateau,  of  Paris,  was  one  of  the  ear- 
liest to  experiment  with  a  steam  turbine  having  a  single  wheel  of 
the  Pelton  type.  The  essential  features  of  his  English  patent  on 
this  are  shown  in  Figs.  36  and  37,  which  represent  the  wheel  vanes. 
He  intended  primarily  to  produce  a  reversible  wheel  and  uses 
buckets  projecting  radially,  with  double  cancave  surfaces,  A,  B, 


SECTION  OF 

PIM    £HO\MNG 

FACE  C  OF  BLAD* 


FACE  OF  //HEEL 


SEGMENT  OF  WHEEL  RIM 
Fig.    36.     Rateau's   Reversing   "Pelton"   Wheel. 


SECTION  OF  RIM 
SHOWING  FACES 
A  AND  B  OF  BLADE 


EARLY  TURBINE  PATENTS 


53 


Fig.  39,  which  form  a  dividing  wedge  at  the  center,  just  as  in  the 
Pelton  water  wheel.  For  reversing  the  direction  of  rotation,  he 
uses  steam  jets  flowing  in  the  opposite  direction  and  impinging 
against  the  backs  of  the  blades.  The  backs  are  shaped  as  shown 
at  C,  with  a  single  concave  surface,  instead  of  with  the  double 
curve,  in  order  to  avoid  any  obstruction  to  the  steam  when  running 
in  the  normal  direction.  While  the  single  curve  is  less  efficient  than 
the  double  curve,  it  answers  the  requirements  for  the  brief  periods 
during  which  the  turbine  has  to  be  reversed. 

When  the  wheel  is  to  be  designed  for  forward  motion  only, 
Rateau  presents  the  construction  of  Fig.  37.     A  and  B  are  two 


Bottom  of  Groove  C 


FACE  OF  WHEEL. 
Fig.  37.     Wheel  for  Forward  Motion  Only. 

concave  vane  surfaces  and  at  the  rear  of  each  bucket  an  inclined 
groove,  C,  is  cut,  represented  by  the  dotted  line,  bp,  in  the  upper 
view.  This  allows  the  jets  of  steam  to  strike  the  buckets,  one  after 
the  other,  without  interference  from  the  successive  buckets  as  they 
come  into  position.  The  sectional  views  at  the  right  are  taken  on 
the  line,  XY,  looking,  in  each  case,  in  the  direction  of  the  arrow 
drawn  under  each  view. 

While  it  is  not  introduced  as  a  definite  claim  in  this  patent,  Pro- 
fessor Rateau  mentions  that  where  the  speed  of  the  fluid  is  too 
great  it  may  be  necessary  to  arrange  these  turbines  in  series  on  the 
same  or  independent  shafts,  in  which -case  the  openings  of  the  noz- 
zles should  all  be  designed  to  deliver  the  same  relative  quantity 
of  fluid  at  the  same  moment.  This  he  would  accomplish  by  using 
distributing  valves  for  supplying  steam  to  the  nozzles,  and  having 


54 


STEAM  TURBINES 


these  vahes  all  operated  positively  from  the  same  source  so  that 
they  would  act  in  unison. 

Parsons,  1895. — In  the  method  of  governing  used  on  Parsons 
turbines,  an  oscillatory  motion  is  given  to  the  throttle  valve  by  an 
eccentric  driven  by  the  turbine,  and  the  extent  of  this  movement 
is  controlled  by  a  governor.  In  the  illustration,  A  is  a  double- 
seated  throttle  attached  to  a  valve  stem,  B,  which  is  connected  with 
a  piston,  C,  working  in  a  cylinder  above  the  valve  chamber.  At  D 


PLAN  OF  LEVERS 


(o)    (o)    CO) 


Fig.    38.     Parson's    Governing    Arrangement. 


is  a  pilot  valve  for  controlling  the  motion  of  the  piston.  The  steam 
enters  the  chamber,  E,  and  flows  downward  through  the  valve  to 
the  turbine.  An  opening  from  E  to  the  space  below  the  piston 
allows  the  steam  to  push  the  piston  upward  against  a  spiral  spring 
which  pushes  it  downward  in  case  the  steam  pressure  underneath 
the  piston  is  relieved.  When  pilot  valve  D  closes  the  port  leading 
from  the  space  below  the  piston,  the  pressure  maintained  under  the 
piston  causes  the  latter  to  rise  and  with  it  the  valve  A,  but  when 
valve  D  uncovers  the  port,  steam  escapes  from  under  the  piston  and 


EARLY  TURBINE  PATENTS  55 

passes  around  to  the  top,  and  together  with  the  spring  serves  to 
close  the  valve. 

A  floating  lever  mechanism  is  used  for  controlling  the  pilot  valve. 
At  F  is  an  eccentric  driven  by  a  worm  and  worm-wheel,  which 
oscillates  lever  G  about  its  fulcrum,  7.  At  point  K,  on  lever  Gf  the 
lever  H  is  fulcrumed.  One  end  of  lever  H  is  connected  to  the 
governor  and  the  other  end  to  the  pilot  valve  D.  The  pilot  valver 
therefore,  is  controlled  both  by  the  motion  of  the  eccentric  and  the 
motion  of  the  governor.  The  eccentric  keeps  the  pilot  valve  and 
hence  the  main  throttle  valve  in  constant  oscillation,  while  the 
movement  of  the  governor  changes  the  positions  of  the  limits  of 
this  motion.  For  example,  if  the  turbine  were  running  with  a 
light  load,  the  valve  would  oscillate  in  the  lower  end  of  its  possible 
path  of  travel  and  would  shut  off  steam  entirely  at  each  oscillation ; 
but  if  the  turbines  were  heavily  loaded,  the  valve  would  be  moved 
upward  by  the  governor  and  its  path  of  travel  would  be  located 
higher,  so  that  steam  would  flow  through  the  valve  continuously, 
although  it  would  be  throttled  more  or  less  as  the  valve  moved  up 
and  down  under  the  action  of  the  eccentric. 

Sebastian  Z.  de  Ferranti,  J#P5. — The  patent  taken  out  by  this 
inventor  is  to  be  classed  with  Hartman's  patent  of  1858  and  that 
of  the  Societe  Anonyme  Maison  Breguet,  1894,  all  three  of  which 
propose  a  compound  turbine  containing  certain  features  employed 
in  the  Curtis  patents  now  used  by  the  General  Electric  Company. 
It  is  to  be  noted,  however,  that,  like  his  predecessors,  Ferranti 
fails  to  specifically  state  that  he  wishes  to  employ  a  diverging  nozzle 
in  combination  with  a  compound  turbine,  which  is  an  important 
feature  of  the  Curtis  type  of  wheel,  although  he  says  that  he  in- 
tends to  utilize  fluid  in  the  wheel,  "after  complete  expansion  and 
the  acquisition  of  the  maximum  velocity,"  which,  under  certain 
conditions,  can  only  be  attained  in  a  diverging  nozzle.  He  advo- 
cates the  use  of  superheated  steam  and  also  refers  to  gas  turbines. 
The  following  is  an  extract  from  his  specifications : 

"I  construct  impact  engines  in  which  the  working  fluid  impinges 
after  complete  expansion  and  acquisition  of  the  maximum  velocity,, 
upon  semi-circular  rotating  blades  fixed  round  the  rim  of  a  motor 
wheel.  The  working  fluid  enters  the  blades  at  a  high  velocity  and 
has  its  direction  reversed,  a  portion  of  its  energy  being  turned  into 


56 


STEAM  TURBINES 


Fig.  39.     One   Plan  for   Compounding. 

work,  and  rotates  the  wheel,  and  then  leaves  at  the  other  side  of 
the  blades  at  a  diminished,  though  still  high  velocity.  I  then  pass 
it  through  a  set  of  standing  semi-circular  blades  of  exactly  the  same 
description  as  the  rotating  blades,  but  with  grooves  in  the  opposite 
direction,  which  reverses  its  direction,  bringing  it  back  to  the 
original  direction  of  motion,  when  it  strikes  the  blades  of  the  second 
wheel  and  delivers  up  a  further  portion  of  its  energy  and  comes 
out  at  a  reduced  velocity.  This  process  is  repeated  until  the 
steam  issues  from  the  last  set  of  blades  with  practically  no  useful 
velocity,  it  having  given  up  nearly  all  its  energy  to  the  rotating 


Fig.  40.     Ferranti's  Turbine. 


EARLY  TURBINE  PATENTS 


57 


blades  of  the  wheel The  object  is  to  convert  the  whole  of 

the  energy  and  pressure  in  the  working  fluid  into  velocity  of  the 
particles,  which  then  react  backward  and  forward  through  the 
rotating  and  standing  blades  of  the  machine,  thus  constituting  an 
impact  multiple  re-active  engine. 

"The  engine  may  be  made  with  one  or  more  expansion  tubes 
according  to  the  power  it  is  desired  to  obtain.  More  or  less  of  these 
expansion  tubes  may  be  used  and  actuated  by  the  governor  accord- 
ing to  the  power  required  for  the  time  being The  expansion 

tubes  stand  tangentially  from  the  periphery  of  the  wheel  and  at  a 


Fig.  41.     A  Second  Plan  for  Compounding. 

slight  angle  to  the  side  of  the  wheel  so  as  to  deliver  its  working 
fluid  in  the  most  suitable  position." 

Fig.  39  shows  the  principle  of  his  scheme,  N  being  the  nozzle,  G 
a  set  of  rotating  vanes,  V  stationary  guide  vanes  reversing 
the  direction  of  motion,  and  so  on.  The  design  of  turbine 
proposed  by  him  is  illustrated  in  Fig.  40,  where  N  is  a  nozzle 
and  G  and  V  the  rotating  and  guide  vanes  respectively,  the 
rotating  vanes  being  attached  to  a  conical  drum  on  the  end 
of  .the  turbine  shaft  and  the  guide  vanes  attached  to  a  casing 
on  the  turbine.  In  this  design  he  plans  to  have  an  increasing 
area  for  the  steam  as  it  flows  through  the  turbine,  after  Parsons' 
plan,  which  is  somewhat  contrary  to  the  statement  of  his 
specifications.  In  Fig.  41  is  still  another  proposed  arrange- 
ment in  which  the  steam,  directed  by  the  nozzle,  N ',  impinges 
against  the  wheel  vanes,  W ,  and  is  then  taken  up  by  .the  U-shaped 
passages,  G^  and  returned  to  the  wheel  vanes,  where  it  is  again 


58 


STEAM  TURBINES 


taken  up  by  the  [/-shaped  passages,  G2.  This  is  another  modifica- 
tion of  the  schemes  advanced  by  Wilson,  Perrigault  &  Farcot  and 
a  number  of  other  early  inventors,  and,  as  previously  stated,  later 
by  Profs.  Riedler  and  Stumpf. 

Curtis,  1896. — The  important  group  of  patents  taken  out  in  this 
year  by  Mr.  C.  G.  Curtis,  inventor  of  the  turbine  manufactured  by 
the  General  Electric  Company,  cover  most  of  the  basic  principles 
of  this  turbine.  The  leading  feature  of  the  Curtis  machine  is  the 
combination  of  a  diverging  or  expanding  nozzle  with  a  compound 


Fig.  42.     Curtis  Turbine. 

turbine  wheel,  although  other  features  are  included  which  had 
been  found  by  the  inventor  to  be  necessary  to  the  successful  opera- 
tion of  a  turbine  of  this  form  of  construction.  As  already  stated 
previous  inventors  have  patented  turbines  in  which  were  combined 
a  nozzle  for  directing  the  steam  against  the  blades  of  the  rotating 
wheels  of  a  compound  turbine,  but  the  Curtis  patents  are  the  first 
to  clearly  claim  the  diverging  nozzle,  as  a  part  of  the  combination, 
and  they  are,  furthermore,  the  first  ones  to  fully  explain  a  practical 
method  for  carrying  out  the  design  so  as  to  make  an  operative  and 
economical  machine.  While  the  requirements  for  the  success- 
ful operation  of  a  compound  impulse  turbine  were  outlined  by 
Moorhouse  in  his  specifications  of  1877,  in  which  he  provided  for 
progressive  expansion  of  the  steam  from  inlet  to  exhaust,  by 


EARLY  TURBINE  PATENTS  59 

using  passages  of  gradually  increasing  areas,  his  patent  was  limited 
to  the  type  of  construction  having  a  succession  of  compartments, 
in  each  of  which  is  a  single  turbine  wheel. 

In  Fig.  42  are  shown  the  elements  of  the  Curtis  turbine.  The 
shaft,  SS,  carries  a  turbine  wheel  on  which  are  two  annular  rows 
of  buckets,  BB.  At  N  is  the  expanding  nozzle  for  directing  the 
steam  against  the  first  ring  of  buckets,  after  which  it  passes  through 
a  group  of  guide  vanes,  G,  to  the  second  ring  of  wheel  buckets. 
In  the  operation  of  a  turbine  of  this  type  the  ideal  condition  would 
be  attained  if  the  expansion  of  the  steam  could  be  complete  in  the 
nozzle  and  then  it  were  to  pass  through  the  turbine  in  virtue  of  its 


Fig.  43.     Preferred  Construction  for  Curtis 
Turbine. 

inertia,  giving  up  a  part  of  its  velocity  to  the  first  ring  of  buckets 
and  the  balance  to  the  second  or  succeeding  rings.  This  would 
allow  the  wheel  to  run  at  a  comparatively  low  velocity,  depending 
upon  the  number  of  times  the  wheel  was  compounded.  That  is,  if 
there  were  a  single  wheel,  as  in  the  De  Laval  type,  it  should  run  at 
approximately  half  the  velocity  of  the  steam ;  but  if  there  were  two 
wheels  instead  they  could  be  run  at  a  lower  velocity  so  that  the 
steam  in  leaving  the  first  wheel  would  have  a  residual  velocity  to  be 
taken  up  by  the  second  wheel.  By  carrying  the  compounding  still 
further  a  still  slower  speed  of  rotation  could  be  used.  Steam, 
however,  is  an  elastic  body,  easily  diffused,  and  has  so  small  a 
mass  that  its  inertia  will  not  carry  it  through  a  succession  of  vanes 
in  the  above  manner,  unless  there  is  an  additional  propelling  force 
generated  to  overcome  the  frictional  and  other  resistances  during 
the  passage  through  the  vanes.  This  is  accomplished  in  Fig.  42  by 
reserving  a  part  of  the  expansion  of  the  fluid  to  take  place  in  the 
guide  passages,  G,  which,  as  shown,  are  made  diverging  for  the 
purpose.  Accordingly,  after  the  steam  leaves  the  first  set  of  vanes 


60 


STEAM  TURBINES 


it  receives  an  additional  impulse  in  the  guide  passages  before  com- 
ing in  contact  with  the  second  set  of  vanes.  The  preferred  con- 
struction, however,  and  the  one  which  is  actually  employed,  is 
shown  in  Fig.  43.  Here  the  steam  is  expanded  in  the  nozzle,  N,  to 
nearly,  but  not  quite  the  final  pressure  of  the  exhaust  pipe,  E.  The 
balance  of  the  expansion  occurs  during  the  passage  between  both 
the  rotating  and  stationary  vanes,  and  the  pressure  within  these 
passages  is,  therefore,  slightly  in  excess  of  the  pressure  within 
the  chamber  in  which  the  wheel  is  rotating. 

The  illustration,  Fig.  44,  is  from  the  so-called  ''stage  patent" 


Fig.  44.     Curtis  "Stage"  Turbine. 

upon  the  Curtis  turbine.  The  cut  shows  each  stage  to  be  com- 
posed of  one  or  more  of  the  compound  elements  that  go  to  make 
up  the  turbine  represented  in  Figs.  42  and  43.  In  each  of  the  two 
stages  the  wheel-and-bucket  arrangement  differs  one  from  the 
other.  In  the  first  casing  are  wheels  A  and  B,  each  carrying  two 
sets  of  rotating  rings  or  vanes,  and  in  the  second  casing  is  a  single 
wheel  with  two  sets  of  blades.  The  advantage  of  dividing  the 
turbine  into  stages  in  this  way,  is  that  there  is  less  leakage  be- 
tween the  guide  vanes  and  the  wheel  vanes,  since  the  differences  of 
pressure  are  less ;  and  there  is  also  less  diffusion  of  the  steam  since 
the  number  of  rows  of  vanes  for  the  steam  to  pass  through  in  each 
stage  is  less  than  would  be  the  case  if  all  the  rows  were  combined 
together  in  one  casing  and  the  steam  were  compelled  to  pass 
through  them  in  virtue  of  the  velocity  acquired  in  the  nozzle  at 
the  beginning. 


EARLY  TURBINE  PATENTS 


61 


A  third  patent  taken  out  in  this  year  deals  with  the  problem  of 
governing,  and  Mr.  Curtis  shows  methods  for  changing  the  quan- 
tity of  steam  supplied  to  the  turbine  without  throttling  the  pressure 
or  reducing  the  velocity  of  flow.  Obviously  an  expanding  nozzle 
of  certain  proportions  is  adapted  only  to  the  steam  pressure  for 


Fig.  45.     Curtis'  Plan  for  Governing  a   Compound  Turbine. 

which  it  was  designed,  and  when  the  pressure  is  throttled  the  noz- 
zle does  not  operate  at  its  highest  efficiency.  It  is  proposed  by 
Mr.  Curtis  to  avoid  .this  by  using  a  nozzle  of  rectangular  cross- 
section  with  one  side  adjustable  in  or  out,  regulating  the  quantity 
of  steam  flowing  through  the  nozzle  without  making  a  great  change 
in  the  ratio  of  the  inlet  and  outlet  areas.  In  Fig.  45  is  a  diagram 
showing  the  principle  proposed,  where  the  turbine  is  divided  into 
two  or  more  stages.  A  is  the  steam  inlet,  terminating  in  a  nozzle 


62 


STEAM  TURBINES 


having  a  sliding  piece,  B}  operated  by  the  rack,  D,  and  pinion,  C. 
This  rack  gears  with  another  pinion,  which  transmits  motion 
through  the  rack,  E,  to  the  pinion,  F,  and  this  in  turn  operates  a 
similar  sliding  piece  in  the  nozzle,  directing  steam  against  the 
second  nozzle.  The  claims  for  the  apparatus  as  used  with  a  com- 
pound turbine  cover  first  the  principle  of  governing  by  changing 
the  volume  without  great  variations  in  the  velocity  of  the  steam 


e  E  E 


Fig.  46.     Design  with   New  Type  of  Expansion 
Nozzle. 

by  means  equivalent  to  the  above;  and  second,  the  simultaneous 
and  proportionate  adjustment  of  the  several  passages  leading  to 
the  turbine  or  connecting  the  different  stages  of  the  turbine. 

Bollmann,  1897. — This  invention  is  an  attempt  to  combine  with 
steam  a  fluid  of  greater  density  so  as  to  reduce  the  velocity  of  the 
jet  impinging  against  the  wheel  vanes.  There  have  been  many 
attempts  to  accomplish  this  result,  some  inventors  preferring  to 
mix  a  heavier  gas  with  the  steam  and  others  to  mix  water  or  some 
less  volatile  liquid,  such  as  oil,  with  the  steam  in  a  manner  similar 


EARLY  TURBINE  PATENTS  63 

to  the  way  steam  and  water  are  mixed  in  a  steam  injector.  One 
of  the  earliest  attempts  to  do  this  was  by  Pelletan  in  1838.  Others 
who  have  proposed  fairly  good  arrangements  of  this  character  are 
Millward  in  1866,  Crumlisk  in  1869,  Miller  and  Collins  in  1896, 
and  Lundell  a  year  after  Bollmann,  in  1898. 

The  chief  interest  in  Bollman's  invention  centers  in  a  new  type 
of  expansion  nozzle  rather  than  in  the  plan  for  using  a  heavier 
working  fluid.  The  nozzle  consists  of  an  annular  slot  between  two 
disks  B  and  C, — the  former  of  which  is  adjustable.  Steam  enters 
through  the  pipe  A  and  flows  radially  out  .through  the  annular 
slot  between  the  disks.  Inasmuch  as  the  diameter  of  this  slot  is 
small  where  the  steam  enters  and  is  larger  where  it  leaves,  the 
steam  expands  in  flowing  through  the  slot,  although  the  faces  of 
the  disks  do  not  diverge.  Escaping  from  the  nozzle,  the  steam 
enters  the  space  D  D  and  there  combines  with  air  and  passes 
radially  outward  to  the  guide  vanes  E  E  E  and  the  wheel  vanes 
F  F  F.  The  plan  is  to  use  this  combined  fluid  in  a  turbine  working 
on  substantially  the  principle  of  the  Curtis  turbine. 

With  this  invention  the  review  of  steam  turbine  patents  will 
close.  The  patents  chosen  for  these  pages,  while  representing  only 
a  part  of  the  best  work  of  inventors  of  the  past  century  in  per- 
fecting the  steam  turbine,  point  the  way  by  which  success  has 
finally  been  attained  and  indicate  the  directions  that  the  paths  of 
progress  in  this  field  will  most  likely  take  in  the  future.  The 
author  believes  that  the  hints  contained  in  the  descriptions  of  these 
inventions  will  prove  of  real  value  to  inventors  who  are  at  work  on 
the  steam  turbine  problem,  as  they  will  give  at  least  a  limited  idea 
of  what  has  already  been  accomplished  and  will  enable  inventors 
to  work  more  intelligently.  In  subsequent  pages  reference  will 
be  made  to  some  of  the  later  inventions  in  connection  with  de- 
scriptions of  the  leading  turbines  now  on  the  market. 

A  WORD  WITH  INVENTORS. 

In  reviewing  the  patents  upon  this  subject,  a  great  many  more 
features  have  been  found  in  the  specifications  which  would  con- 
tribute to  an  unsuccessful  turbine  than  to  a  successful  one.  It 
would  be  out  of  the  question  to  point  out  all  of  these,  but  a  few 


64  STEAM  TURBINES 

of  them  have  appeared  so  frequently  and  have  been  re-invented 
so  many  times  at  the  expense  of  the  inventor  and  to  the  profit  of 
the  patent  lawyer  and  the  United  States  Patent  Office  that  it  will 
be  well  to  give  attention  to  them. 

"Momentum"  Turbines. — One  of  these  is  the  scheme  proposed  in 
the  Bollmann  patent  of  1897.  It  is  improbable  that  any  turbine 
in  which  the  velocity  of  the  steam  is  reduced  by  combining  with 
some  other  fluid  on  the  principle  of  the  injector  can  prove  at  all 
economical  in  its  operation.  The  reason  for  this  is  that  in  com- 
bining the  fluids  by  any  of  the  methods  proposed  the  kinetic  energy 
will  be  reduced ;  and  in  doing  work  upon  the  vanes  of  the  wheels 
it  is  kinetic  energy  which  is  desired.  This  can  be  explained  as 
follows : 

Suppose  steam,  flowing  from  a  nozzle,  to  combine  with  some 
other  fluid,  as  in  the  case  of  the  steam  injector,  where  the  steam 
combines  with  water.  The  steam  imparts  a  certain  velocity  to  the 
water  jet,  or  other  fluid,  if  other  is  used,  and  in  calculating  the 
velocity  it  is  necessary  to  apply  the  principle  of  impact,  that 

Momentum  before  combining=  momentum  after  combining. 
Momentum=massXvelocity ;  and  as 

mass——  — — ,  g  being  acceleration  of  gravity, 

weight 
momen  tum= 5 — X  velocity. 

o 

Applying  the  above  principle,  we  may  calculate  the  velocity  of 
the  combined  jet  as  follows  : — 

Let  V— velocity  of  the  steam. 

F^velocity  of  the  combined  fluid. 

W= weight  of  fluid  (water  or  otherwise)  combined  with 
the  steam. 

Then,  assuming  one  pound  of  steam  to  be  used,  we  have, 


g 


EARLY  TURBINE  PATENTS  65 

Now,  if  we  suppose  all  the  kinetic  energy  of  the  jet  be  used  by 
the  turbine,  the  capacity  of  the  jet  for  doing  work  is  represented  by 


in  the  case  of  the  steam,  and  by 

(W+l) 


in  the  case  of  the  combined  fluid.  Let  us  assume  that  one  pound  of 
water  or  other  fluid  is  used  for  each  pound  of  steam.  It  is  evi- 
dent from  above  formula  (1)  that  the  velocity  of  the  combined  jet 
will  then  be  one  half  the  velocity  of  the  steam,  while  the  weight 
will  be  double. 

The  kinetic  energy  of  the  steam, 

F2 


will  therefore  be  two  times  that  of  the  combined  jet,  which  is 

(W+l)   V* 

H 

This  does  not  mean  that  energy  has  been  lost.  In  this  process 
the  heat  energy  of  the  steam  is  first  converted  into  kinetic  energy, 
giving  the  steam  jet  high  velocity  of  flow  at  the  start  ;  second,  part 
of  the  kinetic  energy  is  'converted  back  again  into  potential  energy 
in  the  form  of  heat  or  pressure,  or  both,  after  the  two  fluids  com- 
bine ;  and  the  kinetic  energy  remaining  is  all  that  is  available  for 
doing  work.  If  the  turbine  is  to  be  efficient,  this  potential  energy 
must  be  utilized  by  converting  it  again  into  kinetic  energy,  and  it 
must  be  acknowledged  that  so  many  transformations  would  entail 
serious  losses,  if,  in  fact,  it  were  possible  to  make  them  at  all. 

A  Misconception  of  Reaction.  —  It  has  been  demonstrated  many 
times  that  the  reaction  of  a  jet  of  water  or  steam,  is  not  altered  by 
holding  an  obstruction  in  the  pathway  of  the  jet,  unless  the  ob- 
struction is  placed  near  enough  to  the  mouth  of  the  nozzle  to  choke 
the  flow.  This  has  not  been  realized  by  some  inventors  who  have 
schemed  on  turbines  similar  to  Fig.  47,  where  steam  enters  the 


66  STEAM  TURBINES 

hollow  radial  arms  through  the  trunnion  A  and  discharges  at  the 
orifices  B  and  C.  Those  inventors  who  have  provided  notches 
such  as  N  N  for  the  steam  to  strike  against  have  not  improved  the 
efficiency  of  the  machine  in  any  way.  Likewise  those  who  have 
arranged  for  two  rotating  elements,  one  of  which  may  be  repre- 
sented by  the  arms  in  Fig.  47,  and  the  other  to  consist  of  a  ring 
rotating  in  the  opposite  direction  and  containing  blades  N  N,  have 
done  nothing  to  improve  the  efficiency.  If  the  blades  N  N  are 
properly  shaped,  as,  for  instance,  in  the  Seger  turbine,  speed  re- 


Fig.   47.    A   Useless   Construction. 

duction  may  be  secured  by  this  means,  but  there  would  theoretically 
be  no  improvement  in  the  efficiency.  Other  inventors  have  at- 
tempted to  produce  turbines,  combining  the  principle  of  the  rotary 
engine,  in  which  the  blades  move  through  closed  compartments 
and  the  steam,  after  impinging  against  the  blades,  reacts  against 
an  abutment.  None  of  these  various  schemes  are  likely  to  be  suc- 
cessful, and  inventors  are  advised  to  adhere  to  the  plan  of  first  pro- 
viding means  for  converting  as  much  of  the  potential  or  heat 
energy  of  the  steam  into  kinetic  energy  as  possible,  and  then  using 
this  energy  to  the  best  possible  advantage  according  to  the  well- 
proven  laws  of  the  hydraulic  turbine. 


CHAPTER  III 


SIMPLE  IMPULSE  TURBINES. 
The  De  Laval  Steam  Turbine. 

History  and  Characteristic  Features. — The  De  Laval  turbine  is 
the  invention  of  Gustaf  De  Laval,  the  famous  Swedish  scientist. 
De  Laval  received  an  advanced  technical  education  as  a  prepara- 
tion for  his  career  and  has  shown  himself  a  versatile  inventor  along 
several  lines  in  his  chosen  profession.  Previous  to  his  turbine 


Fig.  1.     De  Laval  Wheel,  Flexible  Shaft  and  Nozzles. 

achievements  his  most  noteworthy  invention  was  in  connection 
with  the  centrifugal  separators,  which  have  become  of  enormous 
importance  in  the  dairy  industry  of  the  world.  It  was  in  the 
manufacture  of  these  that  he  conceived  the  idea  of  the  steam  tur- 
bine as  the  ideal  motor  for  the  separator,  which  must  itself  run  at 
extremely  high  speeds.  As  recorded  in  Chapter  II.,  he  used  the 
reaction  type  at  first.  In  1888  he  experimented  with  the  diverging 
expansion  nozzle,  which  is  mainly  responsible  for  the  efficiency  of 
the  De  Laval  turbine.  The  high  velocity  of  the  steam  in  the 
diverging  nozzle  necessitated  considerably  higher  wheel  velocities 
than  he  had  used  before,  and  it  was  found  to  be  very  difficult  to 
balance  a  wheel  accurately  enough  to  avoid  destructive  pressure 


68 


STEAM  TURBINES 


upon  the  bearings  under  such  conditions.  He  therefore  adopted 
the  flexible  shaft,  which,  with  the  diverging  nozzle,  is  employed 
in  the  De  Laval  turbine  of  to-day.  These  characteristic  features 
are  represented  in  the  familiar  illustration,  Fig.  1. 

General  Description. — Fig.  2  is  an  external  view  of  an  electric 
generating  set  consisting  of  a  De  Laval  turbine  direct-connected 
to  double,  direct-current,  Bullock  generators.  The  generator  is  at 
the  left,  the  turbine  at  the  extreme  right,  and  between  the  two  are 
the  casings  inclosing  the  speed  reduction  gears  connecting  the 
turbine  and  generator  shafts.  The  turbine  wheel  rotates  within 
a  steel  casing  and  on  one  end  of  its  shaft  is  a  small,  double,  spiral 


Fig.  2.     De  Laval  Electric  Generating  Set. 

pinion,  which,  in  the  smaller  sizes,  meshes  with  a  large,  double 
spiral  gear.  In  the  large  sizes  two  double  gears  are  placed,  one  on 
each  side  of  the  pinion,  which  thus  balances  the  thrust  of  the  trans- 
mission. 

In  the  large  engraving,  Fig.  3,  is  a  horizontal  sectional  view  of  a 
turbine  taken  in  the  plane  of  the  turbine  and  gear  shafts.  Starting 
at  the  right,  W  is  the  turbine  wheel  attached  to  the  flexible  shaft, 
which  latter  is  supported  on  each  side  of  the  wheel  by  bearings  held 
in  the  casing  by  ball  and  socket  joints.  The  pressure  within  the 
turbine  casing  is  practically  atmospheric  pressure  when  running 
non-condensing,  and  is  equal  to  the  pressure  of  the  condenser  when 
running  condensing.  Under  the  latter  conditions,  these  bearings 


70  STEAM  TURBINES 

should  be  tight  to  prevent  leakage  of  air  into  the  casing,  and  they 
must  at  the  same  time  be  able  to  move  slightly,  in  case  of  flexure 
of  the  shaft.  They  are,  therefore,  held  to  their  seats  by  spiral 
springs  N  bearing  against  a  collar  0  made  in  the  form  of  a  socket. 
At  the  other  end  of  the  flexible  shaft  are  the  spiral  pinions  K,  sup- 
ported on  each  side  by  bearings  C  in  the  wheel  casing.  These 
pinions  mesh  with  the  gears  /  /,  as  indicated. 

The  speed  reduction  between  the  pinion  and  gears  is  about  in 
the  ratio  of  10  to  1  for  all  sizes  of  turbines.  The  speeds  of  the 
turbine  wheels  range  from  about  30,000  revolutions  per  minute  for 
a  7-horse-power  to  10,600  for  a  300-horse-power  turbine ;  and  the 
speeds  of  the  large  gears  range  from  about  900  to  3,000  revolutions 
per  minute.  The  peripheral  speed  of  the  turbine  wheels  ranges 
from  about  515  to  1,380  feet  per  second,  while  the  peripheral  speed 
of  the  gears  is  100  feet  per  second  or  slightly  more,  for  all  sizes. 
These  speeds  of  the  gear  shaft  are  found  to  be  well  adapted  to 
driving  generators  and  other  apparatus,  such  as  centrifugal  pumps, 
blowers,  etc.  Such  apparatus  is  driven  through  flexible  couplings 
taking  power  from  the  outer  ends  of  the  gear  shafts.  The  coup- 
lings have  a  series  of  pins  F,  Fig.  3,  securely  driven  into  holes  in 
the  circumference  of  the  driving  disks,  and  on  their  outer  ends 
have  rubber  bushings  E,  which  fit  in  corresponding  holes  in  the 
disk  attached  to  the  shaft  belonging  to  the  generator  or  other 
apparatus.  These  bushings  are  fitted  with  an  internal  steel  bush- 
ing D,  which  slips  over  the  end  of 'pin  F,  to  protect  the  rubber. 
This  brings  the  wear  on  the  outside  of  the  rubber  bushing,  which 
presents  a  greater  area  than  the  inside. 

The  governor,  shown  at  M,  is  of  compact  design  and  is  carried 
by  a  short  shaft  made  a  taper  to  fit  in  the  end  of  one  of  the  gear 
shafts.  The  governor  controls  a  throttle  valve  and  also,  in  case  of 
extreme  increase  in  speed,  opens  a  valve  admitting  air  to  the  wheel 
casing  by  means  of  the  lever  V.  The  friction  of  the  wheel  rotating 
in  the  air  checks  its  speed. 

Nozzles  and  Steam  Chest. — In  considering  the  individual  parts 
of  the  De  Laval  turbine,  the  first  to  be  noted  are  the  nozzles  which 
direct  the  steam  against  the  wheel  buckets.  These  nozzles  are  ar- 
ranged about  the  circumference  of  the  steel  casting  which  serves 
as  the  casing  for  the  turbine  wheel.  The  inner  end  of  this  casting 


SIMPLE  IMPULSE  TURBINES 


71 


has  an  annular  closed  space,  separate  from  the  wheel  chamber, 
which  serves  as  a  steam  chest  for  the  turbine,  as  indicated  in 
Fig.  3.  The  inner  ends  of  the  nozzles  open  into  this  steam  chest, 
as  in  the  sectional  view,  Fig.  4.  Here  A  is  the  steam  chest ;  B,  the 
nozzle ;  D,  the  turbine  wheel,  and  C,  the  valve  for  admitting  steam 
to  the  nozzle.  The  divergence  of  the  nozzles  depends  upon  the 
steam  pressure  to  be  used  and  also  upon  whether  the  turbine  is  to 
run  condensing  or  non-condensing.  If  the  latter,  the  turbine  is 
generally  fitted  with  both  condensing  and  non-condensing  nozzles, 
so  that  in  the  event  of  difficulty  with  the  vacuum  the  machine  can 
be  operated  non-condensing  with  a  greater  degree  of  economy. 
The  nozzles  are  turned  to  gauge  on  their  outside  and  reamed  to  the 


Fig.  4.     Nozzle  and  Valve. 

required  taper  on  the  inside.  Over  600  reamers  of  different  tapers 
are  kept  in  the  tool  room  of  the  American  De  Laval  company  for 
this  purpose.  The  nozzles  are  simply  driven  into  place  in  the 
casing,  but  are  threaded  at  their  inner  ends  to  facilitate  removal  by 
means  of  a  jam  nut.  The  taper  of  the  nozzles  ranges  from  about 
6  to  12  degrees  total  taper,  and  they  are  located  with  their  outlet 
about  y$  inch  from  the  wheel  blades. 

Turbine  Wheel  and  Shaft. — The  turbine  wheels  are  all  made  in 
Sweden,  of  a  special  grade  of  high  carbon  steel.  They  are  shaped 
according  to  theoretical  calculations,  so  as  to  offer  nearly  a  uniform 
resistance  throughout  to  the  forces  acting;  but  they  are  made 
slightly  stronger  near  the  center.  A  short  distance  from  the  pe- 
riphery annular  grooves  are  turned  on  each  side  of  the  wheel, 
making  this  the  weak  section,  which  would  be  ruptured  first  in  case 
of  excessive  rotative  speed.  To  further  guard  against  danger  in 


72 


STEAM  TURBINES 


the  case  of  a  wheel  bursting,  the  steel  casing  is  made  strong  enough 
to  sustain  the  shock  due  to  flying  segments  of  the  wheel ;  and  still 
further,  the  hubs  of  the  wheel  extend  into  circular  openings  in  the 
casing  (Fig.  3),  in  which  the  hubs  ordinarily  run  without  touch- 
ing. But  if  the  wheel  rim  should  burst,  what  would  be  left  of  the 
wheel  would  be  out  of  balance  and  would  cause  the  hubs  to  bear 
against  the  casing  with  great  force  and  thus  slow  down. 

Grooves  of  the  shape  shown  in  Fig.  5  are  drilled  and  milled 
through  the  turbine  rim  in  a  crosswise  direction,  and  in  these  drop- 
forged  steel  buckets  are  fitted.  This  construction  enables  buckets 
to*  be  easily  renewed,  as  is  sometimes  necessary  either  because  of 
wear  or  accident. 


Fig.  5.     Method  of  Inserting  Blades. 

In  the  smaller  size  turbines  the  wheels  are  attached  to  the  flexi- 
ble shafts  by  the  method  indicated  in  Fig.  3.  The  hubs  of  the 
wheels  are  bored  out  and  a  thin  steel  bushing  is  drawn  into  the  hub 
by  a  nut  at  one  end.  The  middle  portion  of  the  bushing  is  bored 
tapering  and  fits  on  a  taper  portion  of  the  shaft,  as  indicated.  This 
taper  is  the  standard  y2  inch  per  foot  used  by  the  De  Laval  com- 
pany. After  forcing  the  bushing  on  the  shaft,  it  is  pinned  into 
place ;  but  the  wheel  can  easily  be  removed  by  loosening  the  nut 
and  sliding  it  off  the  steel  bushing.  The  wheels  for  the  larger  tur- 
bines are  made  as  in  Fig.  6.  Here  the  hub  is  solid  at  the  center,  but 
each  end  of  the  hub  is  recessed  and  the  flexible  shaft  is  made  with 
enlarged  flanged  ends  which  fit  into  the  recesses  and  are  bolted  in 
place.  The  recesses  and  shaft  ends  are  machined  on  a  taper  of 


l/2  inch  per  foot. 


SIMPLE  IMPULSE  TURBINES 


73 


The  pinions  are  cut  directly  on  an  enlarged  portion  of  the  shaft, 
the  flexibility  of  the  shaft  making  an  extremely  accurate  balance 
unnecessary  since  the  wheel  and  shaft  reach  the  critical  speed,  so- 
called,  at  about  %  to  %  of  the  normal  number  of  revolutions  of 
the  wheel,  at  which  point  "settling"  takes  place  and  the  parts  pro- 
ceed to  rotate  about  their  center  of 
gravity  instead  of  about  their 
geometrical  center. 

Gears. — Next  in  importance  to 
the  turbine  wheel,  and  probably  first 
in  importance  in  so  far  as  the  suc- 
cessful operation  of  the  turbine  is 
concerned,  are  the  gears  used  to  re- 
duce the  speed  of  the  turbine  shaft 
to  a  point  where  it  is  practicable  to 
utilize  the  power.  It  was  a  radical 
step  on  the  part  of  De  Laval  when 
he  first  attempted  to  run  gearing  at 
so  high  a  speed  as  these  gears 
operate,  and  it  is  safe  to  say  that 
previous  to  the  time  when  De  Laval 
demonstrated  that  gears  would  run 
at  a  linear  velocity  of  upward  of 
100  feet  per  second,  it  would  not 
have  been  supposed  possible. 

The  pinions  are  made  of  .GO-  or 
.70-point  carbon  steel  and  are  a  part 
of  the  flexible  shaft.  The  gears  are 
of  mild  .20-point  carbon  steel  of  :i 
grade  similar  to  that  used  for  loco- 
motive wheel  tires.  For  turbines  up  to  30  horse-power  the  gears 
are  of  solid  steel ;  but  for  sizes  above  that  they  are  made  with  cast- 
iron  centers  with  rims  of  mild  steel.  The  teeth  are  of  fine  pitch, 
ranging  from  about  .15  inch  in  the  smallest  to  .26  inch  in  the 
largest  sizes.  The  success  at  running  these  gears  at  high  speed  is 
due,  in  part,  to  the  fine  pitch  and  the  spiral  angle  of  the  teeth, 
which  thus  brings  a  large  number  of  teeth  in  mesh  at  one  time, 
making  the  working  pressure  at  each  tooth  very  light,  and  re- 


Fig.   6.     Section  of  Turbine  Wheel 
of  the  Larger  Sizes. 


74  STEAM  TURBINES 

ducing  the  likelihood  of  abrasion.  The  dimensions  of  gears  and 
pinions  for  four  sizes  of  turbines  are  shown  in  the  accompanying 
table : 

PINIONS. 

Outside  Number  Depth 

H.  P.  Diameter.  of  Teeth.  of  Teeth. 

10  1.077  21  .075 

75  1-53  19  -1169 

no  1.82  23  .1169 

300  2.65  31  .1275 

GEARS. 

10  10.  i  208  .075 

75  15-7  208  .1169 

no  18.89  250  .1169 

300  29.29  362  .12750 

Oiling  Arrangements. — In  the  high-speed  bearings  oiling  is 
accomplished  by  having  a  shallow  spiral  groove  turned  in  the 
shell,  which  allows  the  oil  to  reach  every  part  of  the  bearing.  In  a 
100-horse-power  machine  this  groove  is  about  %4  inch  deep  and 
^-inch  pitch.  In  connection  with  the  oiling  arrangements  for 
the  bearings,  reference  should  be  made  to  the  design  of  the  wheel 
casing  in  which  the  bearings  are  located.  This  casing  is  in  two 
halves,  divided  on  a  horizontal  plane,  and  the  upper  edge  of  the 
lower  half  has  an  oil  groove  running  around  it,  as  shown  in 
Fig.  3,  to  catch  any  drip  that  may  work  in  between  the  two  halves. 
The  oil  is  carried  down  into  pockets  in  the  casting,  where  the  ring 
oilers  reach  it,  and  these  pockets  are  piped  to  a  gauge  glass,  to  indi- 
cate the  quantity  of  oil  in  them.  The  oiling  of  the  various  bearings 
is  effected  by  means  of  a  single  sight-feed  lubricator  having  tubes 
leading  to  them. 

The  Governor. — Reference  has  now  been  made  to  most  of  the 
principal  parts  of  the  turbine,  with  the  exceptions  of  the  governor 
and  throttle  valve  which  it  controls.  These  are  shown  in  Figs.  7 
and  8  respectively.  The  governor  is  held  in  the  end  of  one  of  the 
gear  shafts  by  the  taper  plug  K,  Fig.  7,  and  is  made  cylindrical  in 
form,  with  its  outer  shell  B  B  cut  longitudinally  into  two  halves 


SIMPLE  IMPULSE  TURBINES 


75 


which  form  the  governor  weights.  These  weights  are  fulcrumed 
at  A  A  and  have  pins  C  C  which  press  against  a  collar  D  which 
takes  the  thrust  of  the  spiral  springs  located  within  the  governor. 
The  movement  of  the  governor  is  transmitted  through  the  center 
spindle  G  to  the  bell-crank  lever  L,  which  is  balanced  by  a  spiral 
spring  N.  The  shaft  supporting  this  lever  passes  through  the 
valve  casting  on  the  inside  of  which  are  a  pair  of  arms  connecting 


Fig.  7.     Governor  and  Connection  with  Throttle  Valve. 


with  a  double-seated  throttle  valve  as  shown.  In  the  steam  pipe 
above  the  valve  is  a  wire  cylindrical  screen,  to  prevent  any  large 
particles  of  scale  or  other  material  likely  to  damage  the  turbine 
from  passing  through.  It  is  to  be  noted  that  the  connection  be- 
tween the  center  spindle  G  of  the  governor  and  the  bell-crank  lever 
L  is  a  flexible  connection,  and  that  at  the  right  is  a  valve  T,  which 
connects  through  the  passage  P  with  the  wheel  casing.  In  case  the 
throttle  valve  should  stick  and  the  turbine  speed  go  up,  the  gov- 


76 


STEAM  TURBINES 


ernor  would  have  power  enough  to  overcome  the  pressure  of  the 
spring  at  the  connection  H,  and  the  pin  O  would  strike  the  spindle 
of  the  valve  T,  which  latter  would  admit  air  to  the  vacuum  cham- 
ber in  which  the  wheel  revolves.  This  would  immediately  put  an 
air  brake  on  the  wheel  and  prevent  an  acceleration  of  speed.  If 
for  any  cause  the  speed  becomes  excessive  this  action  takes  place. 
In  a  paper  read  before  Society  of  Arts,  Boston,  in  1904,  Charles 


Fig.  8.     Throttle  Valve. 


Garrison  states :  "To  show  the  action  of  the  vacuum  breaker 
more  clearly,  I  started  a  150-horse-power  turbine  with  all  nozzles 
open,  the  nozzles  being  designed  for  150  pounds  gauge  pressure 
and  26  inches  vacuum.  The  condenser  was  shut  down  and  the 
turbine  exhausted  against  the  atmosphere,  and  with  these  condi- 
tions the  turbine  would  not  come  up  to  full  speed  with  no  load." 


SIMPLE  IMPULSE  TURBINES 


Special  Applications  of  the  De  Laval  Turbine. 


77 


The  turbine,  as  built  by  the  several  De  Laval  companies,  by 
C.  A.  Parsons  &  Co.  in  England,  and  Sautter,  Harle  &  Co.  in 
France,  has  been  applied  to  many  special  uses  besides  that  of 
dynamo  driving.  In  the  United  States  more  installations  of  this 
character  have  been  undertaken  by  the  De  Laval  company  than  by 
any  other  manufacturer. 

Application  to  Centrifugal  Pumps. — Until  recently  the  centrif- 


Fig.   9.     Application  to   Compound   Centrifugal   Pump. 

ugal  pump  has  not  been  considered  as  efficient  as  the  plunger 
pump,  mainly  because  of  the  low  speed  and  imperfect  design  of 
such  apparatus;  and  it  has  also  been  adapted  only  to  low  lifts. 
As  the  De  Laval  Steam  Turbine  company,  however,  were  building 
high-speed  turbines,  which  ran  at  almost  exactly  the  speeds  re- 
quired for  maximum  efficiency  in  centrifugal  pumps  adapted  to  the 
different  sizes  of  turbines,  it  was  decided  to  design  a  series  of 
pumps,  of  improved  mechanical  construction,  which  would  admit 
of  the  high  speed  necessary,  and  which  would  also  enable  the 
pumps  to  work  against  high  heads.  The  smaller  units  consist  of  a 


78  STEAM  TURBINES 

single  pump,  but  the  larger  units  have  two  pumps  driven  by  the 
double-gear  turbine,  one  pump  being  connected  to  each  gear  shaft, 
permitting  their  operation  in  parallel  for  low  pressure,  and  in  series 
when  high  lifts  are  desired.  Standard  sets  are  built  in  sizes  from 
7  to  300  horse-power  for  all  heads  up  to  300  feet,  handling  from 
90  to  26,000  gallons  per  minute.  Lately  a  high-pressure  pump  has 
been  developed  with  the  forcing  pump  having  a  runner  of  very 


Fig.    10.     High-pressure   Centrifugal   Pump    Connected   in   Series. 

small  diameter  attached  to  the  turbine  shaft  and  rotating  at  the 
extremely  high  speed  of  that  shaft.  The  centrifugal  force  de- 
veloped by  the  runner  under  these  conditions  makes  it  possible  to 
pump  against  heads  of  600  to  1,000  feet,  and  to  use  the  pump  for 
boiler  feeding.  A  runner  operating  at  such  high  speed,  however, 
would  force  out  the  water  more  rapidly  than  it  could  draw  it  in  by 
suction,  or  under  atmospheric  pressure  only,  and  hence  the  pressure 
pump  is  flooded  with  water  at  sufficient  pressure  to  ensure  an  ade- 
quate supply,  by  a  pump  on  the  geared  shaft. 


SIMPLE  IMPULSE  TURBINES 


79 


Fig.  11.     Application  to  Blowers  of  the  Sirocco  Type. 

Blower  Sets. — Another  field  where  the  turbine  is  well  adapted  is 
for  direct  connection  to  blowers  for  water  pressures  between  4  and 
21  inches.  The  high  velocity  of  the  turbine  makes  it  feasible  to  use 
such  blower  units  for  locations  and  pressures  where  the  positive 
acting  impeller  blowers  have  been  employed.  The  turbine  is  used 
in  connection  with-  blowers  of  the  Sturtevant  and  Sirocco  type. 
When  directly  connected  to  the  blowers,  the  whole  forms  a  com- 
pact unit,  and  eliminates  the  trouble  from  tight  belts  and  heated 
bearings  met  with  in  attempting  to  use  blowers  for  high  pressures. 


CHAPTER  IV 

THE  PELTON  AND  SIMILAR  TYPES. 

Simple  impulse  wheels  of  the  Pelton  type  have  been  experi- 
mented with  extensively  by  Professor  Rateau  of  Paris  and  Pro- 
fessors Riedler  and  Stumpf  of  Berlin ;  although  Rateau  has  now 
abandoned  this  type  in  favor  of  his  multicellular  turbine,  and  the 


Fig.   1.     Wheel  of  Rateau's  Turbine. 

manufacturers  of  the  Riedler-Stumpf  turbines  appear  to  give 
preference  to  compound  turbines  of  later  design.  The  fact,  how- 
ever, that  the  Pelton  water  wheel,  commonly  known  as  the 
"Hurdy-Gurdy"  wheel,  and  its  later  rival  the  Doble  turbine,  have 
been  so  successful  in  American  water-power  plants  where  there  is 
a  high  head  and  a  high  velocity  of  water  jet  at  the  wheel,  makes  it 
seem  probable  that  steam  turbines  similar  in  principle  to  the 
Pelton  water  wheel  will  be  experimented  with  in  this  country  to  a 
considerable  extent. 


THE  PELT  ON  AND  SIMILAR  TYPES 


81 


Rateau's  Simple  Impulse  Wheel. — Fig.  IMS  an  engraving  of 
Rateau's  wheel,  a  line  drawing  of  which  was  shown  in  the  patent 
review  of  the  second  chapter.  In  Fig.  2  are  sectional  drawings, 
showing  the  details  of  the  turbine  itself  as  constructed  from 
Rateau's  designs.  A  number  of  turbines  similar  to  this  were  built, 
the  one  represented  being  direct-connected  to  a  blower. 

The    Riedler-Stumpf    Wheels*    were    invented    by    Professor 


Fig.  2.     Rateau's  Simple  Impulse  Turbine. 

Stumpf  and  developed  with  the  assistance  of  Professor  Riedler. 
A  wheel  for  a  2,000-horse-power  turbine  is  shown  in  Fig.  3  and  the 
nozzle  ring  for  the  same  in  Fig.  4.  In  a  turbine  of  so  large  a  size 
it  was  necessary-to  use  a  complete  ring  of  nozzles,  but  in  machines 
of  smaller  power  one  or  more  segments  containing  nozzles  are  all 
that  are  required.  The  groups  of  nozzles  are  connected  with  a 
central  distributing  chamber  by  means  of  radial  tubes  so  arranged 

•Described  in  a  paper  read  by  Professor  Riedler  in  Germany. 
See  "Machinery"  for  February,  1904. 


82 


STEAM  TURBINES 


that  steam  may  be  admitted  to  one  or  more  of  the  groups  as  de- 
sired. The  buckets  are  cut  into  the  rim  of  the  wheel,  but  the  noz- 
zles are  placed  obliquely  upon  one  another  in  a  ring  which  sur- 
rounds the  wheel.  The  nozzles  are  of  the  De  Laval  type  made  of 
nickel  steel,  but  are  square  in  cross-section.  They  are  produced 
first  in  the  form  of  round  tubes  and  the  diverging  parts  are  then 


Fig.  3.  Fig.  4. 

Wheel  and  Nozzle  Ring  for  Riedler-Stumpf  Turbine. 

drawn  out  square  and  finally  cut  off  obliquely.  The  details  of  the 
wheel  are  indicated  in  Fig.  5,  which  shows  sections  of  the  buckets. 
It  will  be  noted  that  the  buckets  are  so  formed  as  to  overlap  each 
other  something  like  the  shingles  of  a  roof,  instead  of  being  placed 
one  in  front  of  another  as  in  a  Pelton  water  wheel.  They  are  de- 
signed to  reverse  the  steam  jet  through  the  whole  angle  of  180 
degrees.  In  order  to  reduce  the  velocity  of  rotation  below  that 
obtained  in  the  De  Laval  wheel,  Professor  Stumpf  increased  the 
diameter  of  his  wheels  from  six  to  nine  feet,  and  he  also  found 
it  expedient  to  abolish  the  flexible  shaft  by  giving  unusual  atten- 


THE  PELTON  AND  SIMILAR  TYPES 


83 


tion  to  the  balancing  of  the  wheels,  which  are  in  the  form  of  flat 
disks.  Professor  Riedler  states  that  these  disks  may  be  balanced 
50  that  the  center  of  gravity  will  come  within  .004  of  the  geo- 
metrical center;  also,  that  wheels  6^  feet  in  diameter  will  have  a 
factor  of  safety  of  5  when  running  at  3,000  revolutions  per  minute. 
As  a  result  of  careful  tests  he  is  satisfied  that  nickel  steel  disks  of 
this  size  may  be  procured  which  are  practically  free  from  internal 
strains. 


SECTION  C-D 


SECTION  A-B 


Fig.   5.     Sectional  Views  of  Riedler-Stumpf  Wheel. 

A  Stumpf  Patent,  issued  in  1903,  specifies  double  U-buckets 
with  a  dividing  ridge  milled  out  from  the  solid  rim  of  the  wheel  as 
described  above.  Fig.  6  is  from  the  patent  drawing.  He  proposes 
to  use  a  series  of  guide  passages  having  the  general  contour  of  the 
dotted  line  shown  at  A,  which  gather  up  the  steam  after  it  has 
issued  from  the  two  sides  of  the  wheel  and  return  it  in  a  single, 
solid  stream  at  the  center,  where  it  impinges  a  second  time  against 
the  blades  of  the  same  wheel.  In  Fig.  7  is  shown  a  modified  form 
of  the  guides  having  partitions  extending  in  the  direction  in  which 
the  steam  flows,  for  the  purpose  of  insuring  an  equal  division  of 
the  steam  jet  over  the  whole  breadth  of  the  return  buckets,  with  a 
view  to  preventing  choking  of  the  steam  at  the  point  where  it 
strikes  the  wheel  vanes. 


84 


STEAM  TURBINES 


It  is  pointed  out  by  Professor  Stumpf  that  steam,  in  reversing 
its  direction  of  flow  in  a  turbine  wheel,  acquires  sufficient  centrif- 
ugal force  to  increase  its  pressure,  frequently  by  a  considerable 


SECTION    ON  X-Y 

,  Fig.  6.     Subject  of  Stumpf  Patent. 

amount,  so  that  in  leaving  the  vanes  there  is  a  sudden  explosive 
expansion  of  the  steam,  causing  a  scattering  of  the  jets.  By  catch- 
ing the  steam  in  the  return  buckets  as  it  leaves  the  wheel,  and 
bringing  the  streams  together  in  a  solid  jet  again  at  the  center,  he 
aims  to  overcome  this  action  and  to  produce  a  more  efficient  type 
of  compound  turbine.  His  claims  are  broad  ones,  applying  to  the 
combination  of  admission  nozzles,  a  turbine  wheel  with  double 
buckets  and  double  return  buckets. 

In  Figs.  8  and  9  are  two  arrangements  that  have  been  adopted 
for  applying  the  features  of  the  Stumpf  patent.    The  wheel,  instead 


Fig.  7. 


of  having  double  U-buckets,  has  a  single  U-shaped  bucket.  The 
steam  flows  from  the  nozzle  and  strikes  against  one  side  of  the 
buckets,  then  passes  around  to  the  other  side  and  escapes  to  the 
guides,  which  again  change  its  direction  of  flow  and  cause  it  to 


THE  PELT  ON  AND  SIMILAR  TYPES 


85 


)C)o/ 

//  A 


STEAM  ENTERS 


Fig.    8.     Showing   Arrangement   of    Guides. 

impinge  a  second  time  against  the  wheel.    The  course  of  the  steam 
is  indicated  by  the  letters  A  B  C  D  in  the  illustration. 

In  Fig.  9  is  the  arrangement  where  there  are  double  U-buckets. 
The  steam  enters  through  the  nozzle  in  the  direction  of  the  arrow 
A,  the  jet  divides  when  it  strikes  the  buckets,  flowing  in  the  direc- 


8TEAM  ENTERS 


Fig.    9.     Modified   Arrangement   of    Guides. 


86  STEAM  TURBINES 

tion  indicated  by  the  arrows  B  B  and  C  C,  and  finally  impinges  the 
second  time  against  the  buckets  in  the  direction  of  the  arrow  D. 

In  a  subsequent  chapter  further  reference  will  be  made  to  the 
Riedler-Stumpf  turbine,  in  which  the  compound  principle  is  applied 
by  using  two  or  more  wheels  instead  of  a  single  wheel  as  here 
described. 

Claims  Made  for  the  'Pelton  Type. — The  best  presentation  of  the 
claims  for  a  Pelton  type  of  wheel  for  steam  turbine  purposes  has 
been  given  by  John  Richards  in  a  paper  before  the  Technical 
Society  of  the  Pacific  Coast  in  May,  1904.*  He  argues  that  steam 
impulse  wheels  as  usually  built  are  at  fault  because  the  blades  are 
curved  in  one  plane  only,  and  consequently  have  but  one  correct 
position  in  the  jet  throughout  the  whole  arc  of  their  movement ;  and 
furthermore  in  nearly  all  cases  are  cut  out  of  solid  metal,  and  have 
angular  or  imperfect  corners.  Whether  the  principles  enunciated 
will  hold  when  using  an  elastic  fluid  like  steam  instead  of  an  in- 
elastic fluid  like  water  can  only  be  told  by  experiment.  The  ten- 
dency of  the  steam  jet  is  to  break  up  into  spray  and  eddy  cur- 
rents, whereas  a  water  jet  will  "hang  together"  for  a  longer  period. 
This  act  may  have  an  important  bearing  on  the  question  of  the 
spacing  of  the  buckets  in  a  Pelton  type  of  wheel  when  used  with 
steam. 

The  following  is  extracted  from  Mr.  Richards'  paper,  beginning 
with  his  objections  to  the  type  of  bucket  employed  in  impulse  steam 
turbines  as  now  constructed,  such  as  in  the  De  Laval  : 

First.  It  increases  the  weight  and  number  of  the  buckets  about  fivefold 
in  the  attempt  to  secure  impingement  of  the  steam  jets  normal  to  the  straight 
faces  of  the  buckets. 

Second.  It  distorts  the  course  of  reaction  from  a  possible  angle  of  15  de- 
grees to  an  angle  of  20  to  30  degrees  required  to  secure  clearance. 

Third.  It  makes  necessary  a  side  application  of  the  jet,  introducing 
lateral  stress  on  the  wheels  and  inducing  vibration. 

Fourth.  It  augments,  in  proportion  to  the  added  number  of  buckets,  the 
amount  of  fluid  friction.  Not  to  include  the  resistance  of  corners. 

The  number  of  buckets  is  an  important  matter.  It  is  a  sequence  of  the 
angle  of  impingement,  and  this  again  is  a  sequence  of  the  bucket's  shape,  as 
will  be  shown  further  on.  The  surface  or  fluid  friction,  which  offers  a  con- 

*Published  in  the  Journal  of  the  Association  of  Engineering  Societies,  Philadelphia, 
September,  1904. 


THE  P ELTON  AND  SIMILAR  TYPES 


87 


siderable  resistance  and  loss,  is  in  proportion  to  the  number  of  buckets  em- 
ployed, and  should  be  considered  in  this  connection. 

Most  of  the  steam  turbine  buckets  now  made  have  angular  corners,  and, 
when  there  are  not  such  corners,  the  end  walls  of  the  buckets  are  so  distant 
from  the  jet  as  to  lose  reactive  effect  in  that  direction.  We  long  ago 
learned  to  keep  water  out  of  sharp  corners  in  hydraulic  practice. 


Figs.    10    and    11. 

Figure  10  shows  how  the  line  of  impingement  varies  in  respect  to  the 
straight  faces  of  radial  buckets,  and  there  is  no*  way  of  securing  impinge- 
ment even  approximately  normal  to  the  straight  faces,  except  by  employing 
a  large  number  of  buckets  set  close  together.  The  result  is  much  the  same 
whether  the  jets  be  applied  tangentially  or  on  the  side,  as  shown  in  Fig.  11, 
where  the  angle  of  entrance  is  20  degrees  and  that  of  discharge  36  de- 
grees in  conformity  with  the  practice  of  the  De  Laval  company. 


The  trend  of  practice  in  tangential  water  wheels  has  been  to  wider 
spaces  between  the  buckets,  better  angles  for  discharge,  and,  recently,  to 
uniformly  curved  buckets,  as  hereinafter  explained. 

In  Fig.  11  the  entrance  and  discharge  angles  embrace  an  arc  of  56  degrees, 
which,  by  reducing  the  number  of  buckets,  could  be  reduced  to  36  degrees  or 


88  STEAM  TURBINES 

less  if  the  problem  of  oblique  impingement  were  out  of  the  way.  Fig.  12 
shows  spacing  for  tangential  buckets  to  secure  an  easy  discharge  at  20  de- 
grees. 

In  the  Riedler-Stumpf  turbines,  the  angle  of  discharge  is  180  degrees. 
In  other  words,  the  discharge  is  opposite  the  jet,  but  this  calls  for  in- 
creased surface,  more  width  and  weight  for  the  revolving  member,  and 
expensive  work  in  construction,  which  are  hardly  offset  by  countervailing 
advantages,  and  which  certainly  prevent  a  cheap  and  general  manufacture 
of  the  machine. 

Mr.  Richards  then  contends  that  buckets  of  steam  turbines 
should  be  curved  in  all  planes  approximately  as  shown  in  Fig.  13, 
taken  from  a  form  of  water  buckets  of  an  advanced  type  by  W.  A. 


SECTION  ON  xy 
Fig.    13.     Buckets   Suggested  by    Richards. 

Doble  of  San  Francisco.  These  are  of  double  concave  or  cup  form, 
in  order  to  permit  direct  and  balanced  impingement  at  the  various 
angles  in  which  they  are  presented  to  the  jet,  and  have  a  central 
dividing  wedge  to  permit  tangential  application.  The  bucket  is 
notched  at  A,  following  the  construction  of  certain  water  wheels, 
to  permit  the  passage  of  the  jet  beyond  and  through  the  buckets 
as  they  come  into  position,  so  that  it  will  impinge  against  the 
buckets  in  advance  which  have  reached  a  position  where  the  jet 
will  act  upon  them  efficiently.  He  estimates  that  about  one  bucket 
for  each  8  degrees  of  arc  will  be  sufficient  for  wheels  from  20  to  40 
inches  diameter.  This  is  less  than  one-fifth  the  number  now  em- 
ployed for  wheels  having  the  ordinary  type  of  buckets. 

Zoelly's  Patents. — In  1900  a  patent  was  taken  out  by  Heinrich 
Zoelly  for  a  turbine  wheel  of  the  Pelton  type,  but  with  radial  arms, 
the  outer  ends  of  which  serve  as  vanes  for  the  wheel.  These  arms 
decrease  in  cross-section  as  they  approach  the  periphery  of  the 


THE  PELTON  AND  SIMILAR  TYPES 


wheel  and  thus  are  proportioned  to  resist  the  stresses  due  to  centrif- 
ugal force,  which  are  greater  near  the  center.  The  first  claim  of 
this  patent  is  for  "the  combination  in  a  turbine  wheel  of  radial 
buckets  separated  from  each  other  for  a  part  of  their  length,  each 
bucket  having  its  receiving  face  channeled  for  the  greater  portion 
of  its  length,  and  a  pair  of  flat  disks  inclosing  said  buckets  from 
their  inner  ends  for  a  greater  portion  of  the  length  of  the  channeled 
part  of  the  buckets."  In  Fig.  14  A  and  B  are  two  sectional  views 


SECTION  ON  M-N  | | 

Fig.    14.     Wheels    Patented   by   Zoelly. 

of  the  wheel,  C,  an  enlarged  front  view  of  one  of  the  blades,  and 
D  and  E,  enlarged  sections  of  the  blade  on  the  lines  x  y  and  m  n 
respectively. 

At  F  and  G  are  details  of  a  wheel  now  used  in  the  Zoelly  tur- 
bine described  in  a  subsequent  chapter.  In  this  type,  patented  in 
1903,  the  steam  is  directed  against  the  blades  on  one  side  of  the 
wheel  and  escapes  on  the  other  side.  Sections  of  the  blades  are 
shown  at  H.  The  first  claim  is  for  a  "turbine  blade  constructed 
with  a  gradually  increasing  longitudinal  thickness  and  a  longi- 
tudinal cavity  of  substantially  uniform  depth."  Some  of  the  other 


90 


STEAM  TURBINES 


claims  relate  also  to  the  method  of  clamping  the  blades  in  position, 
and  the  use  of  spacing  blocks,  b  b,  between  them. 

Richards'  Patent. — The  idea  presented  in  the  Zoelly  patent  of 
1900  is  carried  a  step  further  in  a  patent  issued  to  J.  Richards  for 
a  wheel  in  accordance  with  his  ideas.  The  wheel  consists  of 
buckets,  B,  light  in  weight  and  drop  forged  on  the  ends  of  radial 
arms,  which  are  attached  to  a  central  nave  by  pins  inserted  between 
the  arms,  as  indicated  in  the  illustration.  The  buckets  are  spaced 
further  apart  than  is  usual  in  turbines,  because  they  are  designed  to 
be  concave  and  reactive  through  a  considerable  angle  of  rotation, 
and  thus  will  absorb  the  energy  of  the  jet  sufficiently  throughout 


Fig.  IS.     Turbine  Wheel  Proposed  by  Richards. 

this  range.  The  arms  of  the  wheels  are  not  covered  by  plates  at 
the  sides,  as  in  the  Zoelly  design ;  the  intention  of  the  inventor 
being  that  the  inside  of  the  turbine  casing  shall  be  machined 
smooth,  and  the  steam  allowed  to  rotate  with  the  wheel  within  the 
casing.  The  claim  is  in  substance  for  a  wheel  having  a  single  hub, 
of  a  diameter  within  the  zone  of  disruptive  centrifugal  strain,  with 
equidistant  radial  sockets  formed  therein ;  strong  radial  stems  fast- 
ened in  the  sockets ;  and  concave  reactive  buckets  integrally  formed 
on  the  extremities  of  the  stems. 

Turbine  Designed  by  John  Richards. — In  Fig.  16  is  a  turbine 
proposed  by  John  Richards  and  patented  by  him  in  1903.  In  this 
turbine  he  uses  a  wheel  like  that  outlined  above.  The  casing  for 
the  wheel  is  finished  smooth  inside  and  the  steam  is  supposed  to 
rotate  with  the  wheel  in  the  casing.  The  gearing  of  transmission 


THE  PELTON  AND  SIMILAR  TYPES 


91 


is  inclosed  with  the  motor  wheel  and  operates  within  the  vapor 
contained  in  the  casing.  The  three  reasons  for  this  are:  (1)  That 
he  believes  the  wheels  will  wear  better  when  steam  lubricated; 
(2)  that  noises,  if  present,  will  be  abolished;  (3)  that  by  this  con- 
struction packing  glands  on  the  spindle  of  the  motor  wheel  may  be 
abolished.  The  journals  for  the  wheel  spindles  are  to  be  hardened 
and  ground,  mounted  in  pivoted  split  shells  of  cast  iron,  and  like 
the  gearing,  exposed  to  the  vapor  of  the  wheel  chamber,  which  is 
at  low  pressure. 


Fig.    16.     Turbine   Patented  by  John  Richards. 

Mr.  Richards  is  an  advocate  of  simple  impulse  wheels  with  gear 
transmission,  because  such  types  are  more  cheaply  constructed 
than  elaborate  compound  turbines.  In  his  paper  before  the  Tech- 
nical Society  of  the  Pacific  Coast,  already  referred  to,  he  gives 
numerous  examples  within  his  experience,  of  high  speed  gearing 
and  rapid  running  machinery,  which  have  given  service  during 
long  periods  of  time. 

Levin's  Turbine. — In  1904  a  turbine  was  patented  by  A.  M. 
Levin  in  which  the  steam  is  expanded  completely  in  a  nozzle  before 
impinging  against  the  buckets  of  the  wheel  and  then  is  used 
several  times  in  succession  upon  the  blades  of  the  same  wheel. 


92 


STEAM  TURBINES 


Steam  is  expanded  in  the  nozzle,  N,  which  projects  it  against  one 
side  of  the  semi-circular  buckets,  B,  of  the  wheel.  The  steam 
passes  around  these  buckets  and  is  projected  outward  against  the 
curved  surface,  C,  of  the  casing  twice  in  succession,  which  at  each 
time  redirects  the  steam  against  the  buckets  of  the  wheel.  The 
curved  surface  of  the  casing  is  stepped  so  that  the  portion  at  C1 
may  be  brought  nearer  to  the  wheel,  and  when  the  steam  reaches 
the  buckets  at  B*  it  is  projected  against  the  surface  at  C1.  As 
drawn,  the  arrangement  is  designed  for  a  wheel  having  a  pe- 
ripheral speed  of  one-tenth  the  initial  velocity  of  the  steam  since  the 
steam  is  projected  against  the  buckets  five  times  in  succession. 
The  buckets  are  semi-circular  in  form. 


SECTION  ON  X-Y 

Fig.  17.     Experimental  Turbine  of  Levin. 

The  steam  proceeds  in  a  succession  of  helical  whirls  after  leav- 
ing the  nozzle,  and  it  is  necessary  that  the  steam  should  be  com- 
pletely expanded  in  the  nozzle  so  that  it  will  be  at  constant  pressure, 
but  have  a  decreasing  velocity  after  leaving  the  nozzle.  The  first 
claim  for  this  wheel  is  for  a  "multiple  impulse  turbine,  comprising 
a  wheel  having  a  row  of  buckets,  an  expansion  nozzle  delivering 
into  said  buckets,  and  a  stationary  reversing  guide  extending  from 
said  nozzle  over  a  number  of  said  buckets,  to  form  a  space  open 
end  to  end  within  which  the  motive  fluid  proceeds  in  a  helical 
whirl  and  is  successively  projected  against  the  buckets  of  said 
wheel." 

Mr.  Levin  built  an  experimental  turbine  on  this  order  which  was 
described  in  "Power"  in  May,  1904.  One  of  the  interesting 
features  is  a  widening  of  the  semi-circular  grooves  constituting  the 
wheel  buckets  and  the  guide  surfaces  at  the  points  where  the  steam 


THE  PELTON  AND  SIMILAR  TYPES 


93 


leaves  the  grooves,  as  clearly  shown  in  Fig.  17  at  D.  This  is  be- 
cause the  steam  becomes  compressed  in  passing  around  the  curved 
surfaces,  and  at  the  points  of  escape,  the  passages  are  widened  to 
allow  the  steam  to  reexpand  to  its  previous  volume.  In  reference 
to  frictional  losses,  Mr.  Levin  states  that  his  tests  indicate  they 
are  not  of  prohibitive  importance,  nor  has  he  found  indications  of 
wear  even  when  moist  steam  has  been  used.  On  the  contrary,  after 
having  run  the  wheel  with  moist  steam,  the  buckets  were  always 
coated  with  a  fine  bluish  film,  apparently  derived  from  oil  carried 
over  from  the  boiler. 


Fig.    18.     Buckets  of    Kerr   Turbine. 

C.  V.  Kerr's  Turbine. — Another  patent  is  that  issued  to  Mr. 
C.  V.  Kerr  for  a  compound  impulse  wheel  of  the  Pelton  type.  The 
buckets  are  made  of  drop  forgings  and  of  such  a  shape  that  they 
may  be  bored  out  perfectly  smooth  by  means  of  a  special  reamer. 
Accordingly,  each  recess  in  the  bucket  on  either  side  of  the  dividing 
wedge  has  a  contour  representing  a  surface  of  revolution. 
Sketches  of  the  buckets  appear  in  Fig.  18.  The  curves  of  the  in- 
terior of  the  bucket  in  a  transverse  direction  must  obviously  be 
circles  in  whole  or  in  part,  as  shown  at  E,  whereas  the  longitudinal 
section  will  show  curves  elliptical  in  shape. 

The  buckets  are  attached  to  steel  disks,  and  in  order  to  withstand 
the  great  strain  he  prefers  to  attach  them,  as  shown  at  A,  by  dove- 


94 


STEAM  TURBINES 


tailing  and  upsetting  the  interlocking  parts,  or  else  by  electric  weld- 
ing. Another  construction  suggested  is  by  riveting,  as  shown  at 
E.  Expansion  nozzles  are  used,  of  the  form  indicated  at  C, 
Fig.  18,  the  tip  of  the  nozzle  being  rather  short  inasmuch  as  the 
turbine  is  divided  into  stages  and  only  a  portion  of  the  pressure  of 
the  steam  has  to  be  reduced  at  each  stage.  Each  nozzle  of  the 
turbine  is  controlled  by  a  hand  valve. 

The  Kerr  turbine  is  being  developed  by  the  Kerr  Turbine 
Company,  Wellsville,  N.  Y.  In  Fig.  19  is  a  diagram  illustrating 
the  arrangement.  Steam  flows  through  a  series  of  nozzles  and 


Fig.  19.     Showing  Principle  of  Kerr  Turbine. 

impinges  against  the  cups  of  the  first  wheel,  which  is  located  in  a 
compartment  by  itself.  It  then  flows  through  another  series  of 
nozzles  and  impinges  against  the  cups  of  a  second  wheel  in  a 
second  compartment,  the  cups  being  enough  larger  to  accommo- 
date the  increased  volume  of  the  steam  at  the  lower  pressure.  It 
again  discharges  into  a  third  chamber,  in  this  case  against  two 
wheels.  This  arrangement  is  followed  throughout,  there  being  an 
increase  either  in  the  size  of  the  cups  or  in  the  number  of  wheels 
as  the  low  pressure  end  is  reached.  This  design  provides  for 
manufacture  in  standard  parts,  because  by  combining  the  units 
turbines  of  widely  varying  powers  can  be  constructed  without 
increasing  the  size  of  the  individual  parts.  The  governor  is  of  the 
throttling  type. 


CHAPTER  V 


COMPOUND  IMPULSE  TURBINES— MULTICELLULAR  TYPE. 
The  Rateau  Steam  Turbine. 

History. — The  practical  work  of  M.  Rateau  on  steam  turbines 
began  in  1894  when  he  constructed  the  first  simple  impulse  tur- 
bine, which  has  already  been  described.  He  soon  abandoned  the 
single- wheel  turbine,  however,  in  favor  of  the  system  with  multi- 


Fig. 


Diagram  Showing  Principle  of  Rateau  Turbine. 


pie  wheels  constructed  on  the  impulse  principle.  His  ideas  shaped 
themselves  about  1897  and  the  following  year  the  firm  of  Sautter, 
Harle  &  Co.,  Paris,  began  the  construction  of  a  900-horse-power 
machine  which  was  experimented  with  until  about  1901,  when  the 
machine  was  brought  to  a  high  degree  of  perfection,  and  since 
which  time  many  others  have  been  built.  A  modified  type  of  the 
Rateau  turbine  is  also  built  under  patents  of  -the  inventor  at 


96 


STEAM  TURBINES 


Oerlikon,  and  the  American  rights  have  been  procured  by  the 
Ball  &  Wood  Engine  Company,  New  York. 

Rateau  Patent. — The  main  features  of  the  Rateau  turbine  are 
covered  by  American  patents  issued  to  Rateau  and  Sautter  in  1903. 
The  turbine  is  of  the  compound  impulse  type,  sometimes  called  the 
multicellular  turbine,  in  which  are  a  number  of  wheels  upon  which 
the  steam  acts  in  succession,  each  wheel  being  in  a  separate  com- 
partment. In  Fig.  1  A  is  the  turbine  casing,  B  B,  etc.,  are  the  ro- 
tating wheels  attached  to  the  shaft,  S,  and  C  C  C  are  diaphragms 
forming  the  separate  compartments.  Steam  enters  through  the 


A 


Fig.     2.     Arrangement 
Guides  in   Rateau 
Turbine. 


of 


intake  pipe  at  D,  passes  between  a  series  of  guide  vanes  at  c,  where 
it  is  directed  against  the  vanes,  b,  of  the  first  wheel.  It  then  passes 
through  the  guide  vanes  in  the  next  diaphragm  and  impinges 
against  the  next  wheel,  and  so  on,  until  the  exhaust  space,  E,  is 
reached.  The  depth  of  the  guide  blades  and  wheel  vanes  increases 
progressively  from  the  inlet  to  the  outlet,  to  allow  for  the  increas- 
ing volume  of  the  steam. 

The  principle  of  this  turbine  as  thus  far  outlined  is  in  no  way 
different  from  that  of  Moorhouse  (Chapter  II.),  but  Rateau  has 
introduced  an  arrangement  of  the  vanes  which  is  new  and  is 
covered  by  this  patent.  In  Fig.  2  the  first  set  of  guide  blades  is 
shown  at  A.  These  are  few  in  number  and  are  arranged  in,  say, 
three  groups  about  the  periphery  of  the  first  disk.  After  passing 


COMPOUND  IMPULSE  TURBINES 


97 


through  the  wheel  the  next  set  of  guide  blades  is  reached  at  B, 
which  consists  of  a  greater  number  to  accommodate  the  increasing 
volume  of  the  steam,  and  these  blades  are  arranged  so  as  to  extend 
by  the  first  set  in  the  direction  in  which  the  steam  flows,  as  shown. 
When  the  steam  reaches  the  first  wheel  it  will  be  carried  along  a 
short  distance  by  the  rotation  of  the  wheel  before  discharging 
into  the  wheel  chamber,  and  a  portion  of  the  next  set  of  guide 
blades  should  be  located  in  advance  of  the  previous  set.  At  C  the 
third  set  laps  by  still  more,  and  finally  a  point  will  be  reached 
where  the  blades  will  extend  around  the  full  periphery'  of  the 
casing.  An  advantage  of  this  arrangement  over  one  in  which 
steam  is  first  admitted  to  the  turbine  around  the  whole  periphery 


ABODE  F  G 

Fig.  3.     Construction  of  Rateau  Diaphragms  and  Wheels. 

is  that  as  the  volume  of  the  steam  at  the  admission  point  is  small 
the  vanes  would  necessarily  have  but  little  radial  depth  at  that 
point  if  they  comprised  a  full  circle,  and  there  would  be  excessive 
friction  of  the  steam  when  flowing  through  them.  In  the  Rateau 
arrangement  the  steam  passages  are  deeper  and  the  volume  of 
steam  passing  is  large,  in  proportion  to  the  rubbing  surfaces  of 
the  vanes.  The  reference  to  this  in  the  claims  of  the  patent  is  as 
follows:  "*  *  *  distributors  arranged  in  the  membranes  to 


B 


100 


STEAM  TURBINES 


direct  the  motive  fluid  directly  upon  the  paddle  blades,  and  said 
distributors  increasing  in  width,  and  overlapping  each  other  suc- 
cessively at  one  end  and  not  at  the  other."  Rateau  also  introduces 
features  of  construction  on  which  claims  are  made,  but  which  are 
in  no  way  tied  with  the  blade  arrangement  mentioned  above. 
Some  of  these  are  shown  in  Fig.  3.  At  A  is  one  of  the  diaphragms 


Fig.  6.     Group  of  Diaphragms  or  Distributors  for  Turbines  of 
Different  Sizes. 

containing  the  guide  vanes.  At  B,  C,  D,  and  E  are  typical  wheels 
consisting  of  steel  disks  either  flanged  around  their  peripheries 
or  else  with  annular  channels  riveted  to  their  peripheries.  In  the 
first  two  instances,  B  and  C,  the  disks  are  dished  to  add  to  their 
lateral  strength  and  in  the  last  two  they  are  flat. 

The  vanes,  which  are  curved  suitably  at  the  points  where  the 
steam  strikes,  are  bent  on  an  angle  and  riveted  to  the  circumfer- 
ence of  the  disk.  At  F  and  G  are  enlarged  details  of  the  vanes,  the 


COMPOUND  IMPULSE  TURBINES 


101 


second  one  showing  a  band  or  shroud  riveted  to  the  outer  circum- 
ference. 

Practical  Notes. — In  designing  the  turbine,  Professor  Rateau 
attempted  to  attain  the  three  following  main  objects:  1.  A  high 
mechanical  efficiency  together  with  as  low  an  angular  velocity  as 


Fig.  7.     Pair  of  Wheel  Disks. 

possible.  2.  A  large  and  at  the  same  time  non-injurious  clearance 
between  the  fixed  and  moving  parts.  3.  The  least  possible  weight 
of  the  whole  machine,  and  especially  of  the  rotating  parts.  In 
Fig.  6  is  shown  a  group  of  diaphragms  or  distributors  for  tur- 
bines of  different  sizes.  The  guide  vanes  are  arranged  in  groups 
about  the  periphery  and  the  number  of  openings  about  each  group 
increases  about  the  exhaust  end  of  the  turbine  until  they  finally 
extend  around  the  whole  periphery.  As  in  an  impulse  turbine  the 


W2  STEAM  TURBINES 

expansion  of  the  steam  is  complete  in  this  distributor,  so  that  the 
steam  acts  upon  the  wheel  in  virtue  of  its  velocity,  and  as  the 
wheel  vanes  are  symmetrical  in  shape,  end  thrusts  are  practically 
eliminated.  The  shaft  passes  through  the  diaphragms  in  bushings 
of  anti-friction  metal.  A  pair  of  wheels  is  shown  in  Fig.  7. 
These  are 'constructed  as  indicated  in  sketch  Fig.-  3. 

The  bearings  of  these  turbines  are  external  and  by  means  of  a 
system  of  spring  packing  are  kept  perfectly  tight.  The  speed  is 
controlled  by  a  centrifugal  governor  acting  by  varying  the  pres- 
sure of  the  steam  delivered  to  the  turbine.  By  means  of  a  by-pass 
in  the  main  steam  pipe  it  is  possible  to  deliver  steam  of  full  pres- 
sure both  to  the  entrance  of  the  turbine  and  to  a  point  in  the  ma- 
chine nearer  to  the  condenser,  thus  enabling  a  higher  power  than 
the  normal  amount  to  be  produced  by  the  machine,  much  in  the 
same  manner  that  a  compound  engine  may  be  used  with  full 
pressure  steam  and  low  pressure  cylinders. 

The  Zoelly  Turbine. 

A  steam  turbine  now  attaining  a  prominent  place  abroad  is  the 
Zoelly  turbine  which  has  been  developed  and  is  now  manufactured 
by  Escher,  Wyss  &  Co.,  Zurich,  Switzerland,  the  famous  manu- 
facturers of  hydraulic  turbines,  water-wheel  governors,  etc.  A 
number  of  large  German  firms,  of  which  the  Krupps  are  one,  are 
reported  also  to  have  formed  a  syndicate  for  the  manufacture  and 
sale  of  this  turbine  on  a  large  scale. 

The  general  arrangement  of  the  Zoelly  turbine  is  evident  from 
the  half-tone  illustration,  Fig.  8,  and  the  line  drawing,  Fig.  9, 
almost  without  explanation.  The  turbine  is  divided  into  two  parts, 
encased  separately,  and  placed  far  enough  apart  to  permit  a  bear- 
ing to  be  located  between  them  for  supporting  the  shaft  at  the 
center.  In  Fig.  8  the  top  of  the  casing  of  the  low-pressure  com- 
partment is  lifted,  exposing  the  wheel  blades  to  view,  and  in 
Fig.  9  the  same  compartment  is  shown  in  section.  The  construc- 
tion of  the  high-pressure  section  is.  entirely  similar  to  the  low-pres- 
sure, except  that  the  steam1  passages  have  less  area. 

There  are  ten  rotating  wheels  constructed  in  the  form  of  circu- 
lar disks,  attached  to  the  same  shaft  and  carrying  curved  blades 
on  their  peripheries.  For  each  wheel  there  is  a  set  of  guide  vanes 


104  STEAM  TURBINES 

for  directing  the  steam  against  the  rotating  blades.  These  guide 
vanes  give  the  steam  the  proper  direction  of  flow  and  allow  it  to 
expand  a  certain  amount  as  it  passes  through  the  guide  passages, 
their  function  being  the  same  as  that  of  the  steam  nozzle  in  the 
De  Laval  turbine  and  the  guide  passages  of  the  Rateau  turbine. 
Each  wheel  rotates  in  a  chamber  by  itself,  the  walls  of  which  are 
formed  by  the  disks  to  which  the  guide  vanes  are  attached.  The 
steam  enters  at  A,  Fig.  9,  and  passes  through  a  throttle  valve 
operated  by  the  governor,  to  the  high-pressure  compartment. 
Here  it  flows  through  the  first  set  of  guide  passages,  and  impinges 
against  the  blades  of  the  first  wheel.  The  guide  passages  permit 
the  steam  to  expand  to  the  somewhat  lower  pressure  of  the  first 
chamber  and  thus  partially  convert  its  potential  energy  to  kinetic 
energy,  which  is  mainly  given  up  to  the  rotating  wheel,  since  the 
steam  leaves  the  wheel  at  a  low  velocity.  The  steam  now  passes 
through  the  passages  of  the  second  guide  disk,  expands  to  a 
lower  pressure  and  is  directed  against  the  blades  of  the  second 
rotating  wheel  in  the  second  compartment,  where  it  again  gives 
up  its  kinetic  energy.  When  the  last  step  is  reached  in  the  low- 
pressure  compartment  the  steam  finally  exhausts,  either  into  a 
condenser  or  into  the  air. 

Guides  and  Wheel  Vanes. — At  the  beginning  of  the  high-pres- 
sure section,  the  guide  vanes  occupy  only  a  part  of  the  periphery 
of  the  turbine,  but  toward  the  end  of  the  low-pressure  part  they 
extend  around  the  whole  circumference.  It  will  be  noted  in 
Fig.  10  that  the  passages  through  the  guide  vanes  have  parallel 
sides;  that  is,  the  walls  do  not  diverge  as  in  the  nozzles  of  the 
De  Laval  turbine.  This  construction  is  based  on  the  well-known 
fact  that  steam  will  expand  and  convert  its  available  heat  energy 
into  kinetic  energy,  or  the  energy  of  motion,  by  flowing  through  a 
nozzle  having  straight,  parallel  sides,  provided  the  final  pressure 
is  not  less  than  .58  of  the  initial  pressure;  whereas,  if  the  final 
pressure  is  less  than  this,  the  walls  of  the  nozzle  must  diverge  in 
order  to  fully  expand  the  steam.  In  the  Zoelly  turbine  the  expan- 
sion occurs  in  successive  steps  and  the  pressure  does  not  drop 
sufficiently  at  any  one  step  to  make  guide  passages  with  diverging 
sides  necessary. 

It  will  be  noted  further,  from  Figs.  9  and  10,  that  the  passages 


106 


STEAM  TURBINES 


through  the  wheel  blades  have  their  inrrer  sides  inclined,  produc- 
ing channels  of  gradually  increasing  area.  This  is  not,  however, 
to  allow  for  expansion.  In  this  turbine,  as  in  others  of  the  impulse 
type,  the  pressure  of  the  steam  does  not  change  in  passing  through 
the  rotating  wheel.  The  pressure  is  uniform  throughout  the 
chamber  in  which  the  wheel  turns,  making  a  drop  in  pressure  in 
passing  through  the  wheel  impossible,  and  hence  the  only  effect  of 


SHAPE  OF  GUIDE  VA-E! 
BEFORE   INSERTING. 


PLAN  OF  GUIDE 
AND  WHEEL  VANES. 


GUIDE  WHEEL 

Fig.  10.     Section  of  Diaphragm  and  Wheel. 

the  sloping  sides  of  the  wheel  passages  is  to  cause  the  steam  to 
flow  smoothly,  without  eddy  currents,  into  the  next  guide  pas- 
sages, which  are  of  larger  area  than  the  ones  preceding.  The  con- 
ditions under  which  the  steam:  flows  through  the  wheel  are  entirely 
different  from  those  which  influence  the  flow  through  the  guide 
vanes,  for  the  latter  have  a  higher  pressure  on  one  side  than  on  the 
other. 

Enlarged  sections  of  the  wheels  and  guides  are  shown  in 
Fig.  10.  The  principal  difference  between  the  Zoelly  turbine  and 
others  of  similar  type,  such  as  the  Rateau,  lies  in  the  construction 
of  the  details  of  the  turbine  wheels.  These  are  designed  to  permit 


COMPOUND  IMPULSE  TURBINES  107 

a  high  rotative  speed  without  straining  the  material,  which  makes 
it  possible  to  reduce  the  number  of  steps  used  in  the  turbine  to  a 
minimum.  The  wheels  are  forged  of  one  piece,  including  the  hub, 
from  Siemens-Martin  steel  and  a  T-slot  is  machined  out  of  the 
circumference  of  the  disk  to  hold  the  blades  and  their  distance 
pieces.  One  side  of  the  slot  is  closed  by  a  wrought  iron  ring,  r, 
which  is  riveted  to  the  disk.  The  blades  are  radial,  and  made  of 
nickel  steel,  highly  polished,  which  resists  erosive  action  very 
effectually.  The  cross-section  of  the  blades  decreases  as  the  radius 
increases,  thus  reducing  the  stresses  due  to  centrifugal  force  to  a 
minimum.  The  blades  and  distance  pieces  are  milled  by  special 
machinery. 

Wheels  and  Disks. — The  disks  which  hold  the  guide  vanes  and 
separate  the  turbine  into  compartments,  are  bored  out  to  receive 
the  hubs  of  the  turbine  wheels.  As  there  is  a  higher  pressure  on 
one  side  of  each  disk  than  on  the  other,  it  must  be  steam  tight  to 
prevent  leakage  and  strong  enough  to  prevent  deflection.  The 
only  place  where  leakage  can  occur  is  at  the  center,  where  the 
bore  of  the  hub  must  be  loose  enough  for  a  running  fit  over  the 
hub  of  the  wheel.  Annular  grooves  are  turned  in  the  bore  to  re- 
duce this  leakage,  on  the  same  plan  that  grooves  are  sometimes 
turned  in  the  surface  of  a  pump  plunger  for  the  same  purpose. 
The  guide  disks  are  divided  on  a  diametral  line,  with  their  top 
halves  bolted  to  the  top  of  the  casing  and  their  lower  halves  to 
the  base  so  that  the  top  casing  and  guides  can  be  lifted  off,  expos- 
ing the  wheels.  The  joints  of  these  disks  are  ground  to  a  close  fit, 
to  prevent  leakage. 

The  casings  for  the  high-  and  low-pressure  parts  of  the  turbine 
are  mounted  on  the  bed  plate  by  brackets  placed  midway  between 
their  ends  to  prevent  distortion  from  expansion,  due  to  the  heat  of 
the  steam.  There  are  also  large  clearance  spaces  between  the 
guide  vanes  and  wheel  blades,  made  possible  by  the  fact  that  there 
is  but  small  tendency  to  leak  at  these  points,  as  explained  above. 
This  reduces  the  likelihood  of  the  blades  being  displaced  suffi- 
ciently, through  the  heat  of  the  steam,  to  cause  rubbing  when  the 
turbine  is  running. 

Governor. — In  Fig.  11  is  a  detail  drawing  of  the  governor, 
which  controls  the  turbine  by  throttling  the  steam.  A  centrifugal 


108 


STEAM  TURBINES 


governor  acts  on  a  relay  valve,  m,  and  connects  one  side  or  the 
other  to  pipes  a  and  b;  a  being  a  pipe  leading  from  a  reservoir  full 
of  a  liquid,  such  as  oil  or  water,  under  pressure  produced  by  a 
rotary  pump;  and  b  a  return  leading  to  the  suction  well  of  the 
pump.  The  two  pipes  e  and  /  connect  each  end  of  the  valve  to 
the  cylinder  g,  which  is  located  on  top  of  the  main  throttle  valve. 


Fig.   11.     Governor  of  Zoelly  Turbine. 

The  moving  part  of  this  valve  is  attached  to  the  same  valve  stem 
as  the  piston  h,  in  the  cylinder  g.  If  the  load  on  the  turbine  is 
decreased,  the  resulting  increase  in  speed  raises  the  governor 
lever  n,  and  valve  m  makes  a  direct  connection  between  a  and  / 
and  e  and  b.  The  liquid,  entering  cylinder  g,  forces  piston  h 
down,  which  closes  throttle  valve  k  a  corresponding  amount,  re- 
ducing the  pressure  of  the  entering  steam.  The  valve  stem  is  pro- 
longed and  attached  to  the  end  of  lever  n,  and  hence  the  down- 
ward movement  of  the  throttle  valve  moves  the  relay  valve  m 
back  to  its  original  position.  During  this  return  movement  of  the 


COMPOUND  IMPULSE  TURBINES 


109 


valve  m  the  lever  n  pivots  about  its  left  hand  end ;  while,  when  the 
lever  was  originally  moved  by  the  governor,  it  pivoted  about  its 
right  hand  end.  Lever  n  is  thus  what  is  called  a  floating  lever, 
the  fulcrum  of  which  is  shifted  from  one  end  to  the  other,  accord- 
ing to  the  conditions.  When  valve  m  has  been  returned  to  its 
original  position,  no  further  movement  of  the  throttle  valve  can 
occur  until  the  speed  of  the  turbine  changes  again. 


CONDENSER 

OR 

EXHAUST 
PRESSURE 


Fig.   12.     Diagram  of  Hamilton-Holzwarth  Turbine. 


Hamilton-Holzwarth  Turbine. 

A  turbine  is  being  developed  by  the  Hooven,  Owens  & 
Rentschler  Company,  Hamilton,  Ohio,  which  is  on  the  plan  of  the 
Rateau  and  Zoelly  turbines,  but  differs  in  details  of  construction. 
In  units  of  750  Kw.  and  upward  the  turbine  is  divided  into  two 
parts,  the  high-  and  the  low-pressure.  Steam  enters  through  a 
separator  and  passes  through  the  main  inlet  valve  and  the  regu- 
lating valve,  all  of  which  are  below  the  bed  plate.  As  the  steam 
flows  through  the  first  set  of  stationary  vanes  it  forms  a  complete 
ring  instead  of  entering  through  a  part  of  the  circumference  as  in 
the  Rateau  turbine.  The  casing  is  divided  into  compartments  with 


no 


STEAM  TURBINES 


one  rotating  wheel  in  each.    Both  the  stationary  and  the  moving 
vanes  gradually  increase  in  height  toward  the  low-pressure  end. 

Fig.  12  shows  the  scheme  of  the  turbine.  This  is  somewhat  mis- 
leading, in  that  the  stationary  nozzles  apparently  increase  in  area 
from  inlet  to  outlet,  a  construction  that  would  not  be  required 
with  the  small  drop  in  pressure  that  occurs  between  the  different 
compartments.  The  areas  do  not  actually  increase,  however,  be- 
cause the  guide  vanes  are  so  shaped  that  they  are  nearer  together 
at  the  outlet  side  than  at  the  inlet  side,  and  to  compensate  for  this 


SECTION  ON  C-D 


Fig.  13.     Construction  of  Diaphragms. 

they  have  to  increase  in  height  somewhat  from  the  inlet  to  the  out- 
let sides. 

Details  of  Construction. — In  Figs.  12,  13,  and  14  are  certain 
details  of  construction  of  the  turbine.  Fig.  13  shows  the  sta- 
tionary discs  which  are  built  up  of  two  side  pieces  riveted  together. 
Each  vane  is  a  separate  piece  held  by  its  projection  at  its  lower 
end,  which  fits  in  an  angular  groove  between  the  two  disks  at  their 
periphery.  The  vanes  are  of  drop-forged  steel  and  are  secured  by 
rivets.  After  they  are  in  position  their  outside  ends  are  ground 
and  a  steel  ring  is  shrunk  on.  In  case  it  were  not  desired  to  extend 


COMPOUND  IMPULSE  TURBINES 


111 


the  vanes  around  the  whole  periphery  the  forms  used  at  A  and  C 
would  be  employed  as  indicated  in  the  section  on  C  D. 

The  construction  of  the  wheel  is  shown  in  Fig.  14.  It  is  as  light 
as  possible  with  cast  steel  hubs,  to  which  are  riveted  conical  disks 
A.  There  is  a  space  between  the  disks  at  their  periphery  in  which 
are  riveted  their  steel  segments  B.  The  vanes  are  attached  to  these 


DETAILS  OF  VANES. 

SECTION    OF    WHEEL. 
'  Fig.    14.     Construction  of  Wheels. 

segments,  as  shown  in  the  sectional  view  at  D.  They  are  formed 
with  lips  extending  downward  and  these  lips  are  enlarged  at  their 
ends  to  fit  into  the  enlarged  bottom  portions  of  cross  channels  in 
the  segments  B.  The  vanes  are  so  designed  that  the  passages 
through  the  walls  have  a  uniform  area  from  beginning  to  end  and 
they  are  made  hollow  to  reduce  their  weight.  On  the  outer  ends 
of  the  vanes  a  thin,  steel  band  is  shrunk  to  give  an  outside  wall  to 
the  steam  channels.  The  vanes  are  milled  on  both  edges  to  give 
correct  angles. 


112  STEAM  TURBINES 

The  Governor  is  of  the  spring-and-weight  type  and  controls  the 
turbine  by  throttling  the  steam.  It  is  constructed  on  a  relay  sys- 
tem and  controls  the  valve  by  moving  a  small  wheel  across  the 
face  of  the  rotating  friction  disk.  This  disk  is  driven  by  a  worm 
and  worm  wheel  from  the  shaft  which  operates  the  governor.  At 
normal  speed  the  small  wheel  is  at  the  center  of  the  disk  and  is 
moved  out  of  contact  with  it  by  means  of  a  cam.  If  the  turbine 
should  speed  up,  however,  the  wheel  would  be  moved  a  short  dis- 
tance to  one  side  of  the  center  and  the  cam  would  also  move  suffi- 
ciently to  allow  the  wheel  to  come  in  contact  with  the  disk  and  be 
rotated  a  number  of  turns,  thus  closing  the  throttle  valve  a  slight 
amount,  whereupon  the  wheel  would  be  returned  to  the  central 
position  again  and  the  cam  would  throw  it  out  of  contact  with  the 
disk. 


CHAPTER  VI 


COMPOUND  IMPULSE  TURBINES  (Continued). 

The  Curtis  Turbine. 

The  Curtis  turbine  is  manufactured  in  this  country  by  the  Gen- 
eral Electric  Company,  the  large  sizes  at  Schenectady,  N.  Y.,  and 
the  small  sizes  at  West  Lynn,  Mass.  The  Curtis  marine  turbine  is 
being  developed  by  a  company  headed  by  Mr.  Curtis,  the  inventor, 
who  is  conducting  extensive  experiments. 


SECOND  WHEEL 


SLIDE  OPERATED 
BY  GOVERNOR 


Fig.    1.     Early   Form  of   Curtis  Turbine. 

Early  Type.— The.  machine  is  represented  in  its  simplest  and 
earliest  form  in  Fig.  1.  It  consists  of  two  rings  of.  curved  buckets 
mounted  upon  disks  revolving  with  the  shaft.  Between  the  two 
revolving  rings  is  a  group  of  curved  blades  in  the  form  of  a  short 
segment  fixed  to  the  interior  of  the  turbine  case.  The  nozzle  is 
of  rectangular  cross  section,  so  designed  that  one  side  of  it  can 
slide  in  or  out  without  materially  altering  the  ratio  between  the 
inlet  and  outlet  areas  of  the  nozzle.  By  this  means  the  quantity  of 
steam  delivered  is  adjusted  to  suit  the  load,  and  it  is  not  necessary 
to  govern  by  throttling.  An  early  turbine  of  substantially  this 


114 


STEAM  TURBINES 


design,  of  150  horse-power,  was  tested  at  Stevens  Institute  of 
Technology,  Hobokeri,  N.  J. 

Stage  Turbine. — In  its  practical  form  the  nozzles  are  smaller 
in  area  than  in  the  experimental  machine  mentioned  and  are  ar- 
ranged in  groups;  but  the  method  of  governing  is  in  effect  the 
same.  One  design  that  has  been  used  is  shown  in  Fig.  2.  Steam 


STEAM  CHEST 


MOVING  BLADES 
STATIONARY  BLADES 
MOVING  BLADES 
STATIONARY  BLADES 
MOVING  BLADES 


MOVING  BLADES 


STATIONARY 
BLADES 


MOVING  BLADES 


STATIONARY 
BLADES 


MOVING  BLADES 


_ 

i 


Fig.  2.     Stage  Turbine.     Three  Rotating  Rings  of  Buckets  in 
Each  Stage. 

enters  through  the  series  of  nozzles,  forming  a  broad  belt  of  steam, 
and  the  quantity  admitted  is  regulated  by  a  series  of  poppet  valves, 
one  for  each  nozzle.  Regulation  is  by  opening  or  closing  these 
valves  automatically,  which  has  the  effect  of  increasing  or  decreas- 
ing the  quantity  of  steam  flowing,  as  may  be  required,  without 
reducing  the  initial  pressure.  The  turbine  in  Fig.  2  is  a  "stage" 
turbine,  with  two  stages  or  elements,  each  consisting  of  three  ro- 
tating sets  of  blades  and  the  necessary  guide  vanes.  Each  element 
is  incased  in  a  separate  compartment  with  its  set  of  nozzles. 


COMPOUND  IMPULSE  TURBINES  (Continued) 


115 


Since  the  earlier  turbines  were  constructed,  it  has  been  found 
that  better  results  can  be  obtained  by  dividing  the  turbine  into  more 
stages  and  using  only  two  rotating  rings  of  blades  in  each  stage, 
and  this  plan  is  now  followed  in  the  larger  sizes.  Fig.  3  shows 
a  half  section  of  a  turbine  on  this  plan.  The  nozzle  and  its  con- 
trolling valve  are  at  the  top,  at  the  right,  below  which  are  the  two 
sets  of  wheel  blades,  and  the  intermediate  set  of  guide  vanes. 


Fig.  3.     Stage  Turbine.     Two  Rotating  Rings  of  Buckets  in  Each  Stage. 

Then  follow  in  succession  the  second,  third  and  fourth  stages,  with 
the  wheel  in  each  stage  separated  from  the  others  by  the  dia- 
phragms. Only  the  initial  nozzles  are  controlled  by  the  governor, 
but  between  the  first  and  second  stages  are  hand-operated  valves, 
so  that,  should  the  pressure  become  too  high  in  the  first  stage, 
steam  may  be  delivered  through  these  valves  to  the  second  stage. 
In  later  machines  automatic  spring-operated  valves  have  taken  the 
place  of  the  hand  valves. 

Method  of  Reducing  Rotative  Speed. — The  theoretical  principles 
of  the  Curtis  turbine  have  already  been  outlined  in  Chapter  I.  and 


116 


STEAM  TURBINES 


in  the  patent  review  of  Chapter  II. ,  to  which  the  reader  is  referred. 
Although  a  modified  form  of  the  De  Laval  expansion  nozzle  is 
used,  the  rotative  speed  of  the  wheel  is  much  lower  than  in  the 


Fig.  4.     The  First  5,000  Kw.  Turbine,  Installed  at  the   Commonwealth 
Station,  Chicago. 

De  Laval  type,  since  two  or  more  rotating  rings  of  blades  are  em- 
ployed to  utilize  the  high  velocity  of  the  steam  after  it  leaves  the 
nozzle.  In  the  De  Laval  turbine  the  single  wheel  should  run  as 
nearly  as  practicable  at  half  the  velocity  of  the  steam  jet  in  order 
to  absorb  its  energy.  But  in  the  Curtis  turbine  the  speed  is  much 


Fig    5.     Sectional  Elevation  of  2,000  Kw.  Curtis  Turbine. 


118  STEAM  TURBINES 

less  than  half  the  velocity  of  the  steam,  and  when  the  steam  issues 
from  the  first  set  of  blades  it  has  a  high  residual  velocity ;  and  this, 
in  turn,  is  taken  up  in  part  by  the  second  rotating  set  of  blades, 
and  so  on.  This  construction  makes  it  possible  to  utilize  the 
energy  of  the  steam  with  a  comparatively  small  number  of  blades. 
For  illustration,  suppose  steam  to  start  with  a  velocity  of  3,000 
feet  a  second;  once  compounding  would  reduce  the  required 
velocity  of  the  wheel  by  two,  or  to  750  feet  per  second  instead  of 
the  1,500  feet  theoretically  necessary  with  a  single  wheel,  and 
three  rotating  sets  of  vanes  would  reduce  the  velocity  to  500  feet  a 
second. 

Curtis  Vertical  Turbines. — The  first  commercial  turbine  built 
by  the  General  Electric  Company  was  a  600  Kw.  unit,  installed 
in  their  power  plant  at  Schenectady  in  1901.  This  machine  was 
built  on  the  lines  advocated  by  Mr.  Curtis,  with  a  horizontal  shaft 
and  two  stages  with  groups  of  wheels  in  separate  casings,  as  in 
Fig.  2.  Since  the  construction  of  this  machine  all  the  turbines  of 
the  500  Kw.  size  and  larger  have  been  built  with  shafts  in  a 
vertical  position,  and  the  generator  placed  directly  over  the  turbine. 
The  total  weight  of  the  revolving  parts  is  borne  by  a  step  bearing 
at  the  foot  of  the  shaft,  and  the  shaft  is  steadied  and  aligned  by 
three  bearings,  one  at  the  top  of  the  generator,  another  near  the 
foot  of  the  shaft,  and  a  third  between  the  generator  and  the  tur- 
bine. The  sectional  view,  Fig.  5,  shows  the  arrangement  clearly. 
The  different  parts  are  lettered  as  follows : 

A,  spring-weighted  governor ;  B,  generator ;  C,  casing  inclosing 
the  three  turbine  wheels ;  D,  step  bearing ;  E,  outlet  to  condenser ; 
a,  upper  steady  bearing;  b,  lower  steady  bearing;  c,  stuffing  box 
with  graphite  packing  rings;  d,  connection  from  governor; 
e,  mechanism  operating  admission  valves ;  /,  by-pass  for  maintain- 
ing correct  pressure  in  second  stage. 

The  considerations  leading  to  the  vertical  design  are  stated  by 
one  of  the  engineers  of  the  company,  as  follows  :* 

The  relative  positions  of  revolving  and  stationary  parts  are  definitely 
fixed  by  the  step-bearing.  The  stationary  part  is  symmetrical,  easily 
machined  and  free  from  distortions  by  heat.  The  shaft-bearings  are  re- 


*W.   L.  R.  Emmet,  in  a  paper  upon  the  "Steam  Turbine  in  Modern  Engineering," 
read  before  the  American  Society  of  Mechanical  Engineers  in  1904. 


COMPOUND  IMPULSE  TURBINES  (Continued) 


119 


lieved  from  all  strain,  and  friction  is  practically  eliminated.  The  shaft  is 
free  from  deflection  and  can  be  made  of  any  size  without  reference  to 
bearings,  which  can  be  placed  where  convenient  and  operated  with  surface 
speeds,  which  would  not  be  practicable  with  the  horizontal  arrangement. 

These  features  make  possible  the  use  of  a  very  short  shaft,  and  conse- 
quently the  longitudinal  spacing  of  moving  and  stationary  parts  is  very 
little  effected  by  temperature  differences.  The  turbine  structure  affords 


Fig.    6.     2,000    Kw.    Curtis   Turbine. 

support  and  foundation  for  the  generator.  The  cost  of  foundations  is  very 
small,  and  the  solidity  and  alignment  of  foundation  is  not  of  vital  im- 
portance. Much  floor  space  is  saved.  All  parts  of  the  machine  are  con- 
veniently accessible.  Failure  of  lubrication  cannot  injure  the  shaft  or  other 
expensive  parts. 

Pressures  and  Velocities  of  the  Steam. — In  the  four-stage  tur- 
bine steam  is  expanded  in  the  admission  nozzles  from  an  initial 


120 


STEAM  TURBINES 


Fig.  7.     Bucket  Segments — the  Upper  One  for  Low  Pressure  and  the 
Lower  One  for  High  Pressure  Sections  of  Turbine. 

pressure  of  150  pounds  to  58^2  pounds,  thereby  attaining  a 
velocity  of  2,000  feet  per  second.  It  acts  upon  the  two  rows  of 
moving  vanes  and  then,  in  passing  through  the  second  set  of  noz- 
zles, is  expanded  to  about  18^  pounds,  again  acquiring  a  velocity 
of  about  2,000  feet  per  second.  It  here  acts  upon  the  second  se- 
ries of  bucket  wheels  and  is  delivered  to  a  third  set  of  nozzles, 
which  expand  it  to  about  3^  pounds,  imparting  to  it  a  velocity 
of  about  1,600  feet  per  second.  After  acting  upon  the  third  set  of 
wheels  the  process  is  repeated  and  the  steam  is  delivered  to  a 
fourth  set  of  nozzles,  which  expand  the  steam  to  about  1  pound 
absolute,  giving  it  a  velocity  of  1,400  feet  per  second,  which  is 
absorbed  by  the  fourth  set  of  wheels,  and  by  them  the  steam  is  de- 
livered to  the  condenser  with  its  energy  practically  all  extracted. 

Speeds  of  Rotation. — The  speeds  at  which  the  various  sizes 
of  Curtis  turbines  (60-cycle)  operate  are  as  follows: 

500  Kw 1,800  revolutions  per  minute. 

1,000    "   1,200 

1,500    "    900 

2,000    "    900 

3,000    "    720 

5,000    "    720 


Fig.  8.     Bucket  Segment  with  Rim  Riveted  on. 


COMPOUND  IMPULSE  TURBINES  (Continued) 


121 


In  the  smaller  turbines  the  peripheral  speed  is  about  400  feet  per 
second,  and  in  the  larger  ones  it  is  reduced  to  325  feet  per  second. 
The  Turbine  Buckets. — The  most  vital  point  in  a  steam  turbine 
ib  the  buckets,  since  they,  and  the  spaces  between  them,  must  be 
shaped  correctly  to  give  the  proper  direction  of  flow  and  the 


Fig.   9.     Bucket   Cutting   Machine. 

highest  mechanical  efficiency,  and  also  to  provide  for  the  progres- 
sive expansion  of  the  steam.  The  buckets  of  the  Curtis  turbine 
are  cut  out  of  the  solid  metal  by  special  bucket  cutting  machines. 
In  the  smaller  sizes  the  blades  are  cut  from  the  disks  comprising 
the  wheels,  and  in  the  larger  sizes  the  buckets  are  cut  from  seg- 
ments of  steel  or  bronze  and  then  bolted  around  the  periphery  of 
the  disks.  In  Figs.  7  and  8  are  shown  bucket  segments,  in  the 


122  STEAM  TURBINES 

first  instance  as  they  appear  after  machining  and  in  the  second 
with  a  rim  of  steel  riveted  on,  closing  the  outer  openings  of  the 
curved  passages  between  the  buckets. 

Buckets  are  also  made  of  drawn  metal,  the  pieces  being  set  in  a 
mould  and  fixed  in  place  by  pouring  molten  bronze  around  them, 
thus  forming  one  of  the  segments.  In  all  these  constructions  the 
buckets  themselves  are  less  in  width  than  the  rim  of  the  segment, 
so  there  is  no  possibility  of  their  coming  in  running  contact  with 
any  of  the  stationary  parts  of  the  machine. 

While  the  process  of  cutting  the  buckets  produces  very  nicely 
finished  work,  it  is  at  best  expensive,  calling  for  special  machines, 
which  have  taken  a  long  time  to  design  and  develop.  In  all  of 
them  a  single-pointed  cutting  tool  is  employed,  the  tool  being  so 
guided  by  the  mechanism  that  its  cutting  edge  will  be  in  correct 
position  for  cutting  effectively  at  all  points  of  the  curve.  In  older 
machines  the  tool  was  given  a  motion  of  rotation  around  the  cir- 
cumference of  a  circle  (approximately,  depending  on  the  shape  of 
the  buckets),  and  as  it  passed  the  bucket  segment  it  would  remove 
a  chip.  The  tool  was  gradually  fed  into  the  work  as  the  cutting  ad- 
vanced. In  the  latest  type  the  tool  is  given  an  oscillating  motion, 
back  and  forth  across  the  face  of  the  segment.  On  the  forward 
stroke  the  tool  advances  for  the  cut  and  on  the  return  withdraws 
for  clearance.  The  machine  of  this  type  is  partly  pneumatic  in  its 
action,  and  is  an  exceedingly  interesting  piece  of  mechanism, 

Step  Bearing. — In  Fig.  10  is  a  sectional  drawing  of  the  step 
bearing.  It  consists  of  two  cast-iron  blocks,  A  and  B,  one  carried 
by  the  end  of  the  shaft  and  the  other  held  firmly  in  a  horizontal 
position  and  so  arranged  that  it  can  be  adjusted  up  and  down  by 
a  powerful  screw,  S.  Both  blocks  are  recessed  to  about  one  half 
their  diameter  as  shown  at  C  and  into  this  recess  oil  is  forced 
through  the  central  bore  D,  with  sufficient  pressure  to  raise  the 
shaft  slightly  and  support  its  weight  on  the  thin  film  of  oil  which 
flows  out  between  the  flat  faces  of  the  two  blocks.  The  lubricant 
flowing  out  fills  the  space  surrounding  these  blocks  and  rises  be- 
tween the  vertical  bearing  and  the  shaft,  to  the  overflow  E,  where 
it  escapes.  The  whole  structure  is  inside  the  base  and  packing  is 
used,  aided  by  a  low  steam  pressure,  to  insure  that  oil  shall  not 
escape  into  the  vacuum  chamber  above.  The  pressure  required  in 


COMPOUND  IMPULSE  TURBINES  (Continued) 


123 


the  step  bearing  of  a  5,000  Kw.  machine  is  about  1,000  pounds 
per  square  inch  and  this  is  maintained  by  an  electrically  driven 
pump  aided  by  an  accumulator,  so  that  if  the  pump  should  tem- 
porarily fail  the  pressure  would  still  be  maintained.  There  have 
been  a  number  of  instances  where  the  oil  pressure  has  failed  and 
the  lowering  of  the  shaft  and  wheels  after  the  flow,  of  lubricant 
had  stopped  was  at  the  rate  of  about  .01  inch  per  minute.  The 


Fig.   10.     Step  Bearing. 

wear  on  the  blocks,  however,  did  not  appear  to  injure  the  wearing 
qualities  of  the  bearing.  In  some  later  machines  water  has  been 
used  instead  of  Oil  to  support  the  step  bearing. 

Small  Turbines. — In  a  paper  read  before  the  American  Street 
Railway  Association  in  1904,  Richard  H.  Rice  outlines  some  of  the 
features  of  the  small  horizontal  turbines  built  by  the  General  Elec- 
tric Company  at  their  West  Lynn  works.  Certain  particulars  of 
these  are  given  in  the  accompanying  table : 


124 


STEAM  TURBINES 


Rated 
Capacity, 
K.  W. 

Speed  of 
Shaft, 
R.  P.  M. 

Condensing 
or  Non- 
condensing. 

No.  of 

Stages. 

Current. 

Poles. 

Voltage. 

1# 

5,000 

Non-cond. 

1 

Dir.  Cur. 

2 

60 

15 

4.000-4,500 

"         »i 

1 

**        k» 

2 

80-125 

25 

3,600 

11         ti 

1 

U                 11 

2 

125-250 

75 

2,400 

N-c.  &  Cond. 

2 

11        11 

4 

125-250 

100 

8.6UO 

Cond. 

3 

Alt.  Cur. 

2 

2.300 

150 

2.0UO 

N-c  &  Cond. 

3&4 

Dir.  Cur. 

4 

125-250 

300 

1,800 

N-c.  &  Cond. 

3&4 

D.  C  &  A.  C. 

4 

250,  500  &  2,300 

The  three  smaller  sizes  have  two  bearings.  The  turbine  wheels 
are  overhung  on  the  end  of  the  shaft  and  the  shaft  is  in  one  piece, 
with  the  turbine  and  armature  both  mounted  on  it.  Beginning 
with  the  75  Kw.  size  and  upward  the  shafts  are  in  two  pieces  and 
the  sets  have  four  bearings. 

In  the  small  sizes  where  the  wheels  are  overhung  the  front  end 
of  the  case  may  be  taken  off  to  obtain  access  to  the  wheels  and 
intermediates,  and  in  the  larger  sizes  where  four  bearings  are 
provided  the  upper  half  of  the  casing  is  removable  for  the  same 
purpose. 

In  the  four-bearing  sets  the  generator  and  turbine  shafts  are 
united  by  a  flexible  coupling  which  permits  some  little  inaccuracy 
in  the  alignment  of  the  two  shafts  without  affecting  the  operation 
of  the  set.  This  coupling  is  a  modification  of  the  Oldham  coupling, 
the  necessary  flexibility  being  secured  by  the  use  of  links  turning 
on  pins. 

The  lJ/£,  15  and  25  Kw.  turbines  are  of  the  single-stage  type, 
having  a  single  group  of  nozzles  and  three  rows  of  moving  buckets. 
The  larger  sizes  are  multi-stage  and  have  only  two  rows  of  mov- 
ing buckets  per  stage. 

The  bearings  used  in  these  turbines  are  supported  on  spheres. 
The  linings  are  made  in  two  parts  and  lubrication  is  effected  by 
forced  feed  from  a  pump  which  is  geared  to  the  main  shaft  of  the 
turbine  and  supplies  oil  at  a  pressure  of  from  three  to  six  pounds 
per  square  inch. 

Governing  Mechanism. — Various  methods  of  governing  have 
been  experimented  with  by  the  General  Electric  Company.  In 
sizes  of  25  Kw.  or  less,  governing  is  effected  by  throttling  the 
steam  pressure  by  the  direct  action  of  a  powerful  centrifugal  gov- 
ernor. In  the  larger  sized  machines,  however,  each  nozzle  or 


COMPOUND  IMPULSE  TURBINES  (Continued)  125 

group  of  nozzles  is  supplied  with  steam  from  a  poppet  valve 
operated  by  means  of  controlling  mechanism  under  the  influence 
of  the  governor.  One  method  adopted  for  larger  units  consists  in 
the  use  of  a  hydraulic  cylinder  with  a  controlling  valve  actuated 
by  the  governor.  A  movement  of  the  controlling  valve,  caused  by 
a  change  in  the  speed,  admits  oil  to  one  side  or  the  other  of  the 
piston  of  this  cylinder  and  a  movement  of  the  cylinder  results, 
through  the  intermediate  mechanism,  in  the  opening  or  closing 
of  corresponding  poppet  valves.  While  the  governor  remains  in 
any  given  position  the  hydraulic  cylinder  is  also  stationary  and  is 
locked  in  its  position  by  confining  the  oil  in  both  ends  of  the  cylin- 
der. A  movement  of  the  governor  produces  a  corresponding 
movement  of  the  hydraulic  piston,  and  when  this  movement  has 
taken  place  the  parts  come  to  rest.  The  motion  of  the  hydraulic 
piston  is  transferred  to  a  shaft  running  parallel  with  the  bank  of 
nozzles  and  on  which  is  a  series  of  cams  that  actuate  the  valves. 
The  2,000  Kw.  turbine,  Fig.  5,  is  controlled  by  a  hydraulic  gear  of 
this  type.  The  hydraulic  cylinder  is  located  in  a  vertical  position 
above  the  nozzle  valves,  at  the  right,  and  its  plunger  moves  the 
cam  shaft  one  way  or  the  other  according  to  the  position  of  the 
pilot  valve.  The  cam  shaft  is  plainly  visible  in  the  engraving. 
On  some  of  the  largest  machines  horizontal  cylinders  have  been 
employed  instead  of  the  vertical,  placed  between  the  turbine  and 
the  generator,  and  with  the  plunger  operating  the  cam  shaft 
through  a  rack  and  pinion. 

Mechanically  Operated  Gear. — This  gear  is  a  development  from 
steam-engine  practice  and  is  used  on  some  of  the  turbines  manu- 
factured at  the  Lynn  plant  of  the  General  Electric  Company. 
Each  nozzle  valve  is  actuated  directly  by  a  pair  of  reciprocating 
pawls,  one  adapted  to  open  the  valve  and  the  other  to  close  it. 
The  several  pairs  of  pawls  are  pivoted  to  a  common  moving  sup- 
port, which  is  oscillated  by  a  rock  shaft  receiving  its  motion  from 
the  turbine  shaft  through  a  worm  and  wormwheel.  At  the  upper 
end  of  the  valve  spindles  are  crossheads,  in  which  are  milled 
notches  or  teeth  for  the  pawls  to  engage,  and  the  engagement  of 
the  pawls  in  these  teeth  is  determined  by  the  angular  position  of 
shield  plates  controlled  by  the  governor.  These  plates  are  set 
progressively,  one  in  advance  of  the  other,  to  obtain  successive 


126  STEAM  TURBINES 

actuation  of  the  valves.  When  more  steam  is  required,  a  shield 
plate  permits  the  proper  pawl  to  engage  the  crosshead  of  its  valve 
and  open  the  valve  on  the  upward  stroke;  while  if  less  steam  is 
required  the  shield  plates  will  be  moved  by  the  governor  to  such 
a  position  that  the  proper  pawl  will  close  its  valve  during  the 
downward  stroke  of  the  rock  shaft. 

Another  type  of  mechanical  gear,  that  has  been  applied  to 
smaller  units,  has  positively  actuated  valves  that  are  always  either 
in  the  full  open  or  entirely  closed  positions.  Each  valve  has  a 
crosshead  and  block  and  is  actuated  by  a  dog  consisting  of  a  small 
eccentric  strap  with  a  projecting  arm  about  six  inches  long,  pro- 
vided with  two  hooks,  one  adapted  to  pull  the  crosshead  block 
toward  the  eccentric  shaft  and  open  the  valve  and  the  other  to 
push  the  block  away  from  the  shaft  and  close  the  valve.  The 
governor  controls  the  engagement  of  the  hooks.  The  governing 
arrangement  of  this  gear  is  very  ingenious  and  sketches  of  the 
mechanism  will  be  found  in  the  1906  report  of  the  turbine  com- 
mittee of  the  National  Electric  Light  Association. 

Electric  Governing. — One  of  the  earliest  methods  used  for  con- 
trolling the  nozzle  valves,  and  which  is  still  employed,  is  an  elec- 
tric system  in  which  the  action  of  the  valves  is  governed  by 
solenoids  or  magnets  through  which  an  electric  current  passes. 
In  Fig.  11  is  a  diagram  showing  the  principle  of  the  arrangement. 
The  governor  at  G  connects  with  the  cylinder  R,  on  the  surface  of 
which  is  a  series  of  contact  points  arranged  spirally,  so  that  as 
the  cylinder  turns  one  way  or  the  other  these  points  come  in  con- 
tact successively  with  corresponding  points  from  which  the  ver- 
tical wires  extend  arid  close  the  circuit  through  these  wires  in  suc- 
cession. 

Referring  to  the  figure:  A  is  the  supply  wire  for  the  current 
and  B  the  return.  The  current  passes  through  the  switch  S, 
which  ordinarily  is  closed,  and  thence  to  the  wire  and  to  the  cylin- 
der. The  vertical  wires  at  the  left  connect  with  the  magnets  be- 
longing to  the  various  sets  of  nozzles,  but  in  this  diagram  the 
horizontal  wires  leading  to  one  set  of  nozzles  only  are  indicated, 
which  accounts  for  several  of  the  vertical  wires  having  no  appar- 
ent connections.  When  the  cylinder  R  is  so  rotated  by  the  gov- 
ernor as  to  bring  two  contact  points  together  the  current  ener- 


COMPOUND  IMPULSE  TURBINES  (Continued) 


127 


gizes  the  corresponding  magnet  at  T,  and  thence  passes  to  the 
return  wire  B.  Should  the  turbine  speed  be  above  normal  the 
governor  arm  drops  and  breaks  the  current  at  the  switch  S  and 
all  of  the  magnets  are  thrown  out  of  action  and  their  valves 
closed,  shutting  off  steam. 

The  nozzle  valves  are  not  operated  directly  by  the  magnets,  but 
through  the  medium  of  small  auxiliary  valves  which  the  magnets 


Fig.  11.     Diagram  of  Electric  Control  for  Nozzle  Valves. 

control  and  which  serve  to  create  a  balanced  or  unbalanced  steam 
pressure  upon  the  faces  of  pistons  attached  to  the  nozzle  valve 
spindles.  One  of  the  nozzle  valves  is  shown  in  Fig.  12.  The 
valve  V  is  constantly  pushed  downward  by  the  spring  St  and  the 
action  of  the  valve  is  governed  by  the  pressure  in  the  space  above 
the  piston  P,  which  is  fastened  to  the  valve  stem  A.  Communica- 
tion between  the  pilot  valve  and  this  space  is  had  by  the  pipe  / 
and  when  the  pilot  valve  is  in  such  a  position  that  this  pipe  con- 
nection is  open  to  the  atmosphere  the  unbalanced  pressure  under 
valve  V  will  be  sufficient  to  raise  the  valve  and  allow  steam  to 
enter  the  nozzle.  When,  however,  the  pilot  valve  is  in  such  a 


128 


STEAM  TURBINES 


position  that  steam  under  pressure  is  admitted  to  the  space  above 
piston  P,  this  pressure,  in  connection  with  the  spring  S,  forces  the 
valve  to  its  seat. 


Fig.  12.    Nozzle  Valve. 


Fig.   13.     Curtis  Turbine  Governor. 


Curtis  Turbine  Governor. — In  Fig.  13  is  an  outline  of  the  gov- 
ernor for  a  500  Kw.  machine.  The  governor  is  supported  by  a 
flange  keyed  directly  to  the  top  of  the  vertical  shaft  of  the  tur- 
bine and  the  whole  supporting  framework  rotates  with  the  shaft. 
W  W  are  the  two  weights  fulcrumed  at  the  points  indicated  and 
as  the  turbine  speeds  up  the  centrifugal  force  of  the  weights  pulls 
the  lever  L  downward  against  the  resistance  of  the  spring  S. 


COMPOUND  IMPULSE  TURBINES  (Continued)  129 

A  second  spring  at  St  is  arranged  so  that  its  tension  can  be  in- 
creased or  diminished  to  change  the  loading  of  the  governor  and 
thus  bring  the  speed  of  the  governor  within  small  limits.  The 
lever  L  connects  with  the  valve  mechanism. 

Riedler-Stumpf  Turbine. 

This   turbine   has   been   developed   and   manufactured   by  the 
Allgemeine  Electricitats  Gesellshaft,  Berlin,  who  have  now  be- 


Fig.    14.     Buckets   and   Guides   of    Compound    Riedler- 
Stumpf  Turbine. 

come  incorporated  with  a  new  Berlin  organization,  known  as  the 
Union  Electric  Company.  The  object  of  the  Union  Company  is  to 
exploit  in  certain  European  countries  important  steam  turbine 
patents,  chiefly  those  of  Professors  Riedler  and  Stumpf ,  controlled 
by  the  Allgemeine  Company ;  and  the  Curtis  patents,  owned  by  the 
General  Electric  Company  in  this  country.  As  a  result  of  this 
organization,  turbines  are  now  being  constructed  by  the  A.  E.  G., 
combining  features  to  be  found  in  both  the  Curtis  and  the  Riedler- 
Stumpf  machines. 

The  Riedler-Stumpf  Compound  Turbines. — The  single  wheel 
Riedler-Stumpf  turbines  have  already  been  described  in  Chap- 
ter IV.  In  the  compound  turbine  of  this  design,  in  which  two  or 
more  wheels,  or  else  two  or  more  rows  of  buckets  on  the  same 


130 


STEAM  TURBINES 


wheel,  are  used,  each  wheel  is  provided  with  semi-circular  buckets. 
The  steam  is  projected  against  one  side  of  the  buckets  of  the  first 
wheel  and  then  as  it  escapes  from  the  opposite  side,  it  is  collected  by 
curved  guides  which  carry  it  around  to  the  next  wheel.  The 
sketch,  Fig.  14,  shows  the  arrangement,  and  it  will  be  noted  that 
here  it  is  not  necessary  to  arrange  the  guides  spirally  in  order  to 


Fig.    15.     Compound   Riedler-Stumpf  Turbine. 

circumvent  the  nozzle  as  was  done  in  the  case  of  compounding 
with  the  single  wheel  as  described  in  Chapter  IV.  The  letters 
A,  B,  C,  D  indicate  the  direction  of  the  flow  of  steam  in  Fig.  14. 
When  a  considerable  speed  reduction  is  desired,  the  turbine  is 
divided  into  stages  with  two  steps  in  each  stage. 

In  Fig.  15  is  a  sectional  drawing  of  a  two-stage  turbine,  similar 
in  its  arrangement  to  the  Curtis  turbine  made  in  this  country.    The 


COMPOUND  IMPULSE  TURBINES  (Continued)  131 


Figs.  16,  17,  and  18.    A.  E.  G.  Turbines. 

upper  wheel  may  be  classed  as  a  high-pressure  wheel  and  the 
lower  a  low-pressure  wheel.  Steam  enters  as  indicated,  and  is 
conducted  by  suitable  passages  from  the  first  to  the  second  wheel. 
It  then  enters  the  annular  passage  E  E,  where  it  comes  in  contact 


132 


STEAM  TURBINES 


with  condensing  water  which  enters  through  openings  in  the  pe- 
riphery of  this  passage.  The  mingled  steam  and  water  then  flow  to 
the  center,  where  they  enter  between  the  two  rotating  disks  at  R. 
The  condensed  steam  and  the  condensing  water  are  here  thrown 
outward  by  centrifugal  force  and  are  discharged  through  openings 
at  M  M.  This  novel  arrangement  of  the  condenser  insures  thor- 
ough intermingling  of  the  water  and  steam  and  produces  a  high 
vacuum.  The  shaft  of  this  turbine  and  generator  instead  of  being 
carried  on  a  step  bearing  at  the  bottom  is  supported  by  a  bearing  at 
Bt  situated  between  the  turbine  and  the  generator. 


Fig.  19.    End  View  of  A.  E.  G.  Turbine. 

The  A.  E.  G.  Turbine.— Figs.  16,  17  and  18  show  three  of  the 
more  recent  designs  of  the  Allgemeine  Company,  in  two  of  which 
the  principle  of  the  Curtis  type  of  wheel  is  utilized.  The  simplest 
design  is  that  of  Fig.  16, a  single  stage  turbine, the  wheels  of  which 
are  made  with  a  double  bucket  rim.  In  Fig.  17  is  a  four-stage  tur- 
bine with  but  a  single  wheel  in  each  stage ;  while  in  Fig.  18  is  a 
combination  of  the  two  previous  types.  In  each  of  these  designs 
the  turbine  wheels  are  supported  by  the  ends  of  the  shaft  which 
passes  through  the  bearings  and  overhangs,  while  the  generator  is 
located  between  the  bearings. 

Features  of  Construction  of  the  A.  E.  G.  Turbine. — The  use  of 
only  two  bearings  for  a  generating  unit  simplifies  the  construction 


134  STEAM  TURBINES 

to  a  marked  degree,  and  with  a  very  stiff  frame  and  a  heavy  shaft, 
gives  satisfaction.  Owing  to  the  stiffness  of  the  frame  it  is  possi- 
ble to  secure  the  casing  to  it  so  that  it  is  possible  to  expand  and 
contract.  The  wheel  casings  are  of  cast  iron  provided  with  relief 
valves  as  a  protection  against  a  possible  rise  of  pressure.  The 
wheel  is  held  on  the  shaft  end  by  a  flange  and  is  machined  out  of 
a  solid  nickel  steel  disk.  The  steam  inlet  to  the  main  cut-off  valve 
and  to  the  steam  distributing  chest  is  through  a  fine-meshed  screen. 
The  steam  from  the  distributing  chest  is  delivered  to  the  nozzles 
by  a  number  of  pipes  shown  in  end  view,  Fig.  19.  The  bearings 
are  supplied  with  oil  under  pressure  and  are  lined  with  white 
metal.  The  governor  is  fitted  direct  on  the  free  end  of  the  shaft 
and  is  of  the  spring  type.  It  is  placed  inside  the  steam  distribution 
chest  and  openings  which  lead  from  the  latter  to  the  nozzles  are 
closed  or  opened  by  a  steel  band. 


CHAPTER  VII 
REACTION  TURBINES. 

Parsons  Turbines. 

History. — In  1884  the  first  compound  steam  turbine  was  built 
in  England  by  the  Hon.  C.  A.  Parsons.  It  developed  10  horse- 
power at  1,800  revolutions  per  minute  and  ran  for  several  years 
at  Gateshead-on-Thames,  England.  It  consisted  of  two  groups 
of  15  turbines  each,  the  steam  entering  between  them  and  passing 
in  opposite  directions  through  each  group.  In  1888  several  turbo- 
alternators  of  120  horse-power  were  constructed  and  in  1892 


Guides 
Moving  Vanes 


Guides 
Moving  Vanes 


Fig.    1.     Vanes  of  a   Parsons  Turbine. 

the  compound  turbine  was  first  adapted  to  work  with  a  condenser. 
The  first  condensing  turbine  was  of  200  horse-power  and  ran 
with  steam  of  100  pounds  pressure,  slightly  superheated,  and  a 
vacuum  of  28  inches  of  mercury.  A  steam  consumption  equiv- 
alent to  about  16  pounds  of  steam  per  indicated  horse-power 
was  obtained.  This  event  marked  the  beginning  of  the  rapid 
introduction  of  the  turbine,  since  it  demonstrated  its  possibilities 
on  the  score  of  economy.  In  1895  the  manufacturing  rights  of 
the  Parsons  turbine  were  acquired  in  this  country  by  the  West- 
inghouse  Machine  Company,  who  introduced  their  product  com- 
mercially in  1899. 


136  STEAM  TURBINES 

General  Principles. — In  the  Parsons  turbine  there  are  alternate 
rows  of  stationary  guide  vanes  and  moving  wheel  vanes  as  in 
Fig.  1.  The  steam  flows  through  a  fixed  ring  of  directing 
blades,  which  serve  the  purpose  of  steam  nozzles,  onto  a  revolving 
ring  of  similar  blades  and  so  on,  the  pressure  being  reduced  a 
small  amount  at  each  step.  The  revolving  rings  of  blades  act 
both  in  the  capacity  of  buckets  and  nozzles  as  in  any  reaction 
turbine. 

Assume,  for  illustration,  that  the  steam  expands  from  115 
pounds  absolute  to  atmospheric  pressure  in  its  passage  through 
the  turbine  and  that  there  are  40  rows  of  guides  and  vanes  giv- 
ing an  average  drop  in  pressure  of  2^  pounds  at  each  wheel.  If 
steam  were  to  flow  through  an  expanding  nozzle  from  115 
pounds  to  15  pounds  absolute,  its  velocity  would  be  about  2,700 
feet  per  second ;  but,  by  stepping  down  the  pressure  and  allowing 
it  to  expand  an  average  of  23/2  pounds  at  each  stage,  the  velocity 
of  flow  corresponding  to  the  differences  of  pressure  would  be 
only  about  400  feet  per  second.  If  the  Parsons  turbine  were 
purely  a  reaction  wheel,  the  wheel  would  travel  nearly  as  fast  as 
the  steam  when  it  left  the  moving  vanes,  and  in  the  above  illustra- 
tion would  have  a  peripheral  speed  of  nearly  400  feet  per  second. 
In  the  actual  turbine  the  average  speed  is  much  less  than  this, 
requiring  more  rows  of  blades  and  an  immense  number  of  blades. 
In  a  400  Kw.  turbine  there  are  58  rows  of  guide  vanes  and  wheel 
vanes,  or  116  rows  in  all,  aggregating  about  30,000  blades. 

The  pressure  differences  at  each  element  or  set  of  blades  grad- 
ually decrease  from  inlet  to  exhaust,  instead  of  running  uniform 
as  in  the  above  example,  since,  for  a  given  difference  of  pressure, 
the  velocity  of  steam  is  much  greater  at  low  than  at  high  pres- 
sures. Thus,  in  flowing  from  165  pounds  to  155  pounds  absolute, 
a  difference  of  10  pounds,  the  velocity  acquired  is  only  about 
520  feet  per  second ;  while  at  atmospheric  pressure  practically  the 
same  velocity  is  acquired  by  a  drop  of  only  one  pound  in  pres- 
sure or  %0  as  much  as  in  the  first  case.  The  steam  velocities 
are  kept  within  150  feet  per  second  as  a  minimum  at  the  high- 
pressure  end,  and  600  feet  per  second  as  a  maximum  at  the  low- 
pressure  end. 


REACTION  TURBINES 


137 


Westinghouse-Parsons  Turbines. 

Description  of  Parts. — The  elemental  parts  of  a  Parsons  turbine 
are  the  rotor  or  rotating  element,  the  stator,  comprising  the  casing 
and  guide  vanes,  and  the  balancing  pistons.  These  are  shown  in 
Fig.  2,  which  represents  a  Westinghouse-Parsons  turbine.  Steam 
enters  the  chamber  B  at  boiler  pressure  through  the  steam  pipe  A 
and  passes  to  the  right  through  the  first  group  of  blades  which 
gradually  increase  in  height  (see  Fig.  11)  to  chamber  C.  Here, 
to  avoid  excessively  long  blades  as  well  as  many  sizes  of  blades,  it 


Fig.  2.     Sectional  Elevation  of  Westinghouse-Parsons  Turbine. 

is  necessary  to  jump  to  a  larger  diameter  and  the  steam  flows 
through  a  second  set  to  D  and  finally  through  a  third  set  to  space 
E.  The  balancing  pistons  a,  b  and  c  are  of  such  a  diameter  that 
the  steam  pressure  against  them  exactly  balances  the  axial  thrust 
in  the  direction  of  the  steam  flow.  This  thrust  is  composed  of 
three  factors  :  (1)  The  static  pressure  on  the  end  of  the  drum ;  (2) 
the  forward  thrust  on  the  blades  due  to  the  impact  of  the  steam ; 
and  (3)  the  backward  thrust  due  to  the  reaction  of  the  steam  in 
leaving  the  blades.  The  net  result  is  a  forward  thrust.  The  diame- 
ters of  the  pistons  are  approximately  equal  to  the  mean  diameters 
of  the  steam  areas  of  the  different  steps.  The  pipe  F  connects  the 
space  back  of  the  balancing  pistons  with  the  exhaust  chamber.  G 
is  a  coil  for  cooling  the  oil  circulating  through  the  bearings. 


138 


STEAM  TURBINES 


In  Fig.  3  is  shown  a  400  Kw.  turbine  open  for  inspection.  The 
casing  is  made  in  halves  divided  longitudinally  so  that  the  upper 
half  can  be  removed,  exposing  the  rotor,  which  may  then  be  raised 
from  its  bearings,  after  the  bearing  caps  are  removed.  The 
interior  walls  of  the  casing  contain  the  stationary  radial  blades 
corresponding  to  those  on  the  rotating  cylinder.  Starting  at  the 
left,  it  will  be  seen  that  there  are  several  rows  of  blades,  all 
of  the  same  height ;  then  there  is  a  change  to  blades  of  a  slightly 


Fig.  3.     400  Kw.  Westinghouse-Parsons  Turbine  with  Casing  Removed. 

greater  height  and  there  are  several  rows  of  this  size,  and  so  on. 
When  the  mechanical  limit  is  reached  for  size  of  blade,  the 
rotor  is  then  increased  in  diameter,  giving  a  greater  circumference 
and  allowing  shorter  blades.  The  correct  method,  theoretically, 
would  be  for  each  row  to  be  a  little  higher  than  the  previous  one 
throughout  the  turbine.  Practically,  this  is  neither  convenient 
nor  necessary. 

A  more  detailed  description  will  now  be  given  of  certain  parts 
of  the  turbine,  as  made  by  the  Westinghouse  Machine  Company. 

Turbine  Blades. — These  were  formerly  made  of  a  special  cold- 
drawn  bronze,  but  at  the  present  time  drawn  steel  is  extensively 
used.  The  blades  are  secured  in  annular  rings  turned  on  the  out- 


REACTION  TURBINES 


139 


side  of  the  rotor  and  recessed  on  the  inside  of  the  casing,  by  calk- 
ing after  they  are  in  place.  The  blading  material  is  drawn  out 
into  long  strips  and  sawed  up  to  the  proper  blade  length.  The 
blades  are  separated  by  soft  steel  distance  pieces  which  just  fill  the 
grooves  between  them  and  maintain  the  proper  entrance  angle. 
For  the  extreme  low-pressure  blading  of  the  large  turbines  drop- 
forged  steel  blades  are  employed,  with  separators  forged  directly 
on  the  base  of  the  blades.  For  maintaining  uniform  spacing 
between  the  outer  ends  of  the  larger  blades  and  also  to  stiffen  them, 
a  flat  steel  lace  is  threaded  through  openings  near  the  outer  edge 


Fig.  4.     Upper  Half  of  Casing  of  400  Kw.  Turbine. 

of  the  blades,  and  twisted  between  the  adjacent  blades.  The 
smaller  blades  have  j^-inch  clearance  sidewise  and  the  larger  from 
Y-2.  to  1  inch,  so  that  dangers  of  colliding  are  remote;  and  if  this 
occurs  the  turbine  may  still  be  operated. 

Bearings. — When  the  Parsons  turbine  was  first  experimented 
with  trial  shafts  were  run  with  bearings  of  different  descriptions 
up  to  very  high  velocities,  and  it  was  found  that  no  difficulty  was 
experienced,  provided  the  bearings  were  designed  to  have  a 
certain  amount  of  give  or  elasticity.  It  was  then  determined 
that  elasticity  combined  with  frictional  resistance  to  the  trans- 
verse motion  of  the  bearings  gave  the  best  results  and  led  to 
the  adoption  of  the  bearings  shown  in  Chapter  II.  in  connection 
with  the  Parsons  patents.  The  shell  was  surrounded  by  a  series 


140 


STEAM  TURBINES 


of  friction  rings,  each  being  alternately  larger  and  smaller  than 
the  adjacent  ring,  the  small  series  fitting  the  shell  on  the  outside 
and  the  large  series  fitting  the  hole  in  the  bearing  block. 

In  the  Westinghouse-Parsons  turbine  the  bearings  are  made 
up  of  several  concentric  sleeves  instead  of  the  rings  loosely 
fitted  in  the  pedestals.  Oil  circulates  between  the  sleeves,  and 
the  capillary  action  forms  a  fluid  cushion  about  the  several  sleeves, 
which  restrains  vibration  and  at  the  same  time  gives  sufficient 
flexibility  to  allow  the  shaft  to  revolve  about  its  axis  of  gravity 
instead  of  its  geometrical  axis.  The  bearing  proper  is  a  gun 


Fig.   5.     Ribbed  Disk  for  Water-packed   Gland. 

metal  bushing   which   is   prevented   from  turning  by  a  loosely 
fitted  dowel.     Outside  of  this  are  three  other  concentric  tubes. 

Water-Packed  Glands. — In  any  turbine  it  is  necessary  to  pro- 
vide glands  at  the  ends  of  the  casing  to  prevent  the  escape  of 
steam  or  the  admission  of  air  around  the  shaft,  which  latter  is 
detrimental  in  the  case  of  high  vacuum.  Steam-packed  glands 
have  been  used  with  success,  but  the  Westinghouse  company  now 
use  water-packed  glands.  In  the  casing  is  an  annular  groove 
around  the  shaft,  in  which  rotates  freely  a  disk  attached  to  the 
shaft.  The  disk  has  vanes  on  its  faces,  like  the  blades  of  a 
blower.  (See  Fig.  5.)  The  compartment  is  filled  with  water, 
and  when  the  turbine  is  running  the  water  is  thrown  outward 
and  completely  fills  the  outer  part  of  the  annular  space  and 
prevents  the  air  or  steam  from  passing  the  periphery  of  the  disk. 


REACTION  TURBINES  141 

A  similar  device  has  been  employed  in  the  glands  of  centrifugal 
pumps.* 

Lubrication — A  small  pump,  driven  by  a  worm  and  wormwheei 
upon  the  shaft  circulates  oil  through  a  closed  system,  comprising 
in  the  order  of  arrangements,  pump,  oil  cooler,  bearings  and 
reservoir.  The  oil  is  supplied  to  the  bearings  at  the  top,  at  one 
end  where  it  follows  a  groove  in  the  top  of  the  shell,  from  which 
it  is  distributed  around  the  shaft.  A  forced  circulation  under 
high  pressure  is  not  found  to  be  necessary.  The  object  is  to 
maintain  an  oil  film  around  the  journals  so  they  will  never  actually 
come  in  contact  with  the  bearings. 

Governing  Arrangement. 

Description  of  Governor. — Fig.  6  shows  the  governing  mechan- 
ism of  the  Westinghouse-Parsons  turbine.  It  is  substantially 
the  same  as  that  illustrated  in  the  second  chapter  in  connection 
with  the  Parsons  patent  of  1896.  The  governor  is  of  the  centrif- 
ugal type  with  bell-crank  levers,  the  vertical  arms  of  which  carry 
the  balls,  and  the  horizontal  arms  bear  against  the  spiral  spring, 
which  resists  the  centrifugal  force  of  the  balls.  The  tension  of 
the  spring  may  be  adjusted  for  the  purpose  of  synchronizing  two 
alternating  current  generators  when  running  in  parallel.  The 
main  admission  valve  is  actuated  by  the  piston  B,  which  is  con- 
trolled by  the  pilot  valve  A.  Steam  is  admitted  below  the  piston 
through  the  annular  clearance  around  the  main  valve  stem  and 
raises  the  piston  against  the  pressure  of  the  spring.  When  the 
pilot  valve  A,  however,  uncovers  one  of  the  ports  the  steam 

*A  simple  device  for  packing  the  shaft  of  centrifugal  pumps 
in  a  frictionless  manner  was  introduced  some  years  since  by 
Messrs.  Robinson  Brothers  &  Co.,  of  Melbourne,  Australia, 
and  described  in  a  recent  issue  of  London  Engineering.  It 
is  applicable  as  well  to  the  packing  of  the  shafts  of  steam 
turbines,  where  the  efficiency  is  impaired  by  the  leakage  of 
air  into  the  vacuum  end.  The  idea  is  ingenious  in  its  very 
simplicity.  The  bearing  is  cased  in  with  an  annular  cham- 
ber F  which  is  filled  with  water.  Rotary  motion  is  given  to 
the  water  by  means  of  a  disk  or  set  of  vanes  E  attached  to 
the  shaft.  As  the  disk  is  tight  upon  the  shaft,  any  air  to 
reach  the  interior  must  pass  around  the  tops  of  the  blades. 
The  water  in  which  the  tips  are  immersed,  however,  is  under 
greater  pressure,  due  to  centrifugal  force,  than  the  pressure 
of  the  atmosphere  can  overcome,  and  thus  the  air  is  effect- 
ually excluded  without  the  entrance  or  expenditure  of  water  while  the  shaft  is  left 
entirely  free. — Power,  April,  1904. 


142 


STEAM  TURBINES 


Fig.    6.     Governing    Arrangement. 

escapes  from  the  space  under  the  piston,  through  the  small  ex- 
haust pipe  and  allows  the  spring  to  close  the  valve. 

The  pilot  valve  is  governed  both  by  the  motion  of  the  governor 
and  the  reciprocating  motion  of  the  rod  C  which  is  actuated  by 
an  eccentric  driven  through  a  worm  and  wormwheel  from  the 
main  shaft  of  the  turbine.  D  and  E  are  fixed  fulcrums  and  F 
is  a  floating  fulcrum  moving  up  and  down  with  the  governor 
sleeve.  The  reciprocating  motion  of  the  rod  C  is  communicated 
to  the  pilot  valve  A  and  thence  to  the  main  valve,  admitting 
steam  to  the  turbine  in  puffs;  while  the  distance  that  the  valve 


When.Running  Light  Load 

Fig.  7.     Indicator  Diagram  Showing  Effect  of  Reciprocating 
Valve. 


REACTION  TURBINES 


143 


opens  is  determined  by  the  position  of  the  governor.  The  pilot 
valve  was  also  originally  designed  to  act  as  a  safety  stop,  but  an 
auxiliary  and  entirely  separate  safety  stop  is  now  used  for  actu- 
ating a  quick-closing  throttle  valve.  Fig.  7  shows  an  indi- 
cator card  taken  on  a  Parsons  turbine  at  two  different  loads, 
the  indicator  being  attached  to  the  admission  space  A  in  Fig.  2. 
The  indicator  barrel  was  revolved  at  a  constant  speed.  At  light 
loads  the  valve  opens  for  very  short  periods  and  remains  closed 


Fig.  8.     Valve  for  Admitting  High-Pressure  Steam  to  Low-Pressure  End  of 

Turbine. 

for  the  greater  part  of  the  interval.  As  the  load  increases  the 
valve  remains  open  longer,  until  finally  almost  continuous  pressure 
is  maintained  in  the  high-pressure  end  of  the  turbine.  The  net 
effect  of  this  method  of  governing  is  the  same  as  though  steam 
were  throttled  in  the  usual  manner.  While  it  is  true  that  by  ad- 
mitting steam  in  puffs  high-pressure  steam  is  used  at  all  loads  at 
each  oscillation  of  the  valve,  as  has  been  claimed  for  the  device, 
there  is  also  excessive  throttling  of  the  pressure  at  each  puff. 
This  arrangement  has  the  advantage,  however,  of  removing  the 
possibility  of  the  valve  sticking,  since  it  has  a  continuous  motion. 
Auxiliary  Governing  Valve. — A  secondary  admission  valve  is 


144 


STEAM  TURBINES 


provided  to  admit  high-pressure  steam  to  the  second  drum  of  the 
turbine  on  overloads  and  increase  its  capacity  up  to  50  per  cent  or 
more  in  excess  of  the  normal  rating.  This  arrangement  has  the 
further  advantage  that  it  enables  much  better  economy  to  be  main- 
tained under  normal  loading  than  when  the  primary  admission 
valve  only  is  used  for  governing.  In  Figs.  8  and  9  are  diagrams 
showing  the  principle  of  the  valve,  and  its  connections,  taken 
from  the  patent  records,  and  the  sectional  view,  Fig.  11,  shows 
the  actual  arrangement  of  both  primary  and  secondary  valves. 

In  Fig.  8  A  is  the  steam  inlet  and  B  is  a  valve  admitting  high- 
pressure  steam,  when  the  valve  is  raised,  to  the  port  C,  which 


Fig.  9.     Showing   Connections  for  By-Pass. 

connects  with  the  intermediate  part  of  the  turbine.  This  is 
shown  in  Fig.  9,  where  steam  enters  at  A;  B  is  the  valve,  and  C 
is  the  pipe  leading  to  the  turbine. 

The  valve,  B,  is  hollow,  allowing  the  steam  to  pass  through 
to  the  space,  D,  where  it  bears  against  the  under  side  of  the  piston, 
E.  F  is  a  small  passage  leading  from  space,  D,  to  the  space,  H, 
so  that  under  ordinary  conditions  there  will  be  a  balanced 
pressure  on  the  piston  E,  and  the  valve  will  be  kept  seated  by  the 
spring,  vS.  Connecting  with  the  space,  H,  in  which  the  spring  is 
located,  is  a  pipe,  P,  leading  to  a  by-pass  in  the  base  of  the  gov- 
ernor, shown  at  the  left,  which  is  opened  or  closed  by  a  pilot 
valve,  M,  under  control  of  the  governor.  Under  normal  conditions 
the  pilot  valve  will  be  in  the  position  shown,  closing  the  by-pass 
and  preventing  the  escape  of  steam  from  the  chamber,  H.  Should 
the  speed  of  the  engine  decrease  beyond  a  fixed  point,  however, 


146  STEAM  TURBINES 

the  governor  balls  would  move  inward,  which  would  depress  the 
governor  yoke  and  the  pilot  valve,  M,  causing  the  pilot  to 
open  and  allow  steam  to  escape  from  space,  H,  into  the  atmos- 
phere. The  result  would  be  an  unbalanced  pressure  on  the  piston, 
E,  causing  it  to  raise  and  compress  the  spring  and  open  the  valve, 
B,  allowing  the  high-pressure  steam  to  enter  the  low-pressure 
part  of  the  turbine  and  increase  the  power. 

Sectional  View  of  Westinghouse-P  arsons  Turbine. — In  Fig.  11 
on  the  opposite  page  is  a  sectional  view  showing  the  essential 
parts  of  a  turbine,  which  are  lettered  to  correspond  with  the 
following  list  of  parts : 

.9. — Steam  admission. 

V . — Admission  valve  to  high-pressure  end.  This  valve  is  controlled  by 
the  governor  (connections  not  shown)  and  is  oscillated  by  an  eccentric 
driven  by  the  worm  and  wormwheel  at  right-hand  end  of  main  shaft. 

V  g. — Auxiliary  valve  also  controlled  by  governor  and  opened  auto- 
matically in  case  of  overload. 

P ',  P,  P. — Pistons  or  disks  against  which  steam  pressure  acts  to  balance 
the  thrust. 

E,  E,  E. — Equalizing  pipes.  The  two  upper  ones  maintain  steam  pres- 
sures against  the  front  faces  of  balance  pistons  equal  to  the  pressures  in  the 
steps  or  stages  of  the  turbine  having  corresponding  diameters — the  lower 
one  maintains  vacuum  pressure  at  the  back  of  the  large  piston. 

T. — Adjustment  bearing  for  maintaining  the  exact  running  position  of 
the  rotor  and  for  taking  up  any  unbalanced  thrust  not  provided  for  by  the 
pistons.  This  bearing  is  adjusted  endwise  to  locate  the  rotor  in  correct 
position  relative  to  the  stationary  part. 

R.— Relief  valve. 

B. — Exhaust  passage. 

The  arrangement  of  the  governor  has  previously  been  described. 

5,500  Kw.  Turbines. — The  general  view,  Fig.  10,  shows  a 
turbine  of  this  power  of  the  type  that  is  to  be  used  for  the  initial 
equipment  of  the  Pennsylvania  Railroad  terminal  property  in  New 
York  City,  furnishing  electric  power  for  the  trains  passing  through 
the  tunnel  approaches  to  New  York  City  now  in  construction.  The 
space  occupied  by  this  machine  is  approximately  27  feet  8  inches 
by  13  feet  3  inches,  and  the  height  to  the  top  of  the  rail  is  12 
feet.  It  occupies  less  than  %0  square  foot  per  electric  horse- 
power and  develops  20.2  horse-power  per  square  foot  of  floor 
area.  For  the  complete  unit,  including  the  generator,  a  space  of 
the  above  width  and  47  feet  4  inches  long  is  required. 


I  U 


ti- 


148 STEAM  TURBINES 

The  machine  runs  at  750  revolutions  per  minute.  The  con- 
struction is  substantially  that  used  in  the  smaller  machines.  To 
support  the  drum  a  central  steel  quill  is  employed.  Hollow  forged 
steel  ends  are  forced  into  the  ends  of  the  quill  and  constitute  the 
journals.  High-pressure  steam  is  conveyed  to  all  parts  of  the 
quill  structure  to  eliminate  distortion  due  to  expansion. 

The  bearings  of  these  larger  machines  are  of  the  self -aligning 
type,  similar  to  those  employed  in  generators  and  cross-compound 
engines.  The  departure  from  the  oil  cushion  journals  used  in  the 
smaller  machines  is  made  possible  by  the  low  speed  of  rotation. 

Turbines  with  Separate  High-  and  Low-Pressure  Cylinders. — 
Three  turbines,  aggregating  3,750  Kw.,  have  been  installed  in  the 
immense  power  plant  of  the  Interborough  Rapid  Transit  Company, 
New  York  City,  to  furnish  current  for  lighting  the  New  York 
subway.  The  most  striking  feature  of  these  is  the  separation  of 
the  high-pressure  and  low-pressure  sections,  which  provides  for 
a  central  bearing  for  the  turbine  shaft,  thus  reducing  the  distance 
between  bearings  and  allowing  a  much  lighter  shaft.  As  originally 
designed,  these  turbines  were  to  be  equipped  with  reheaters,  placed 
between  the  cylinders,  but  the  reheaters  were  finally  omitted. 
It  was  expected  that  the  drying  of  the  steam  and  possibly  its 
superheating  in  the  receiver,  before  the  steam  entered  the  low- 
pressure  cylinder,  would  effect  an  improvement  in  economy.  Ex- 
periments at  the  Westinghouse  shops,  however,  have  demonstrated 
that  such  is  not  the  case.  Francis  Hodgkinson  states  that  "ex- 
haustive tests  have  shown  the  reheater  to  be  of  little,  if  any,  value 
in  increasing  the  economy  of  the  turbine  when  the  high-pressure 
steam  condensed  in  the  reheater  coils  was  charged  up  against  the 
turbine.  An  improvement  in  the  separator  resulted  in  an  im- 
provement in  the  operation  of  the  reheater,  but,  notwithstanding 
this,  no  advantage  due  to  the  reheater  could  be  observed  and  its 
application  does  not  seem  to  be  warranted,  on  account  of  the 
decreased  compactness  of  the  machine." 

Westinghouse-Parsons  turbines  are  no  longer  built  with  two 
cylinders,  the  whole  tendency  being  toward  compactness.  The 
1,500  Kw.  unit  with  inclosed  generator,  Fig.  12,  is  of  the  latest 
type.  The  objects  of  encasing  the  generator  are  to  reduce  the 
noise  which  sometimes  is  enough  of  a  roar  to  be  disagreeable,  and 


150 


STEAM  TURBINES 


to  enable  a  current  of  air  to  be  forced  through  the  generator  for 
the  purpose  of  maintaining  a  moderate  temperature  and  enabling 
heavier  overloads  to  be  carried  without  danger  of  overheating. 

The  Brown-Boveri  Turbine. 

Description. — In  Fig.  13  is  shown  one  of  the  Parsons  turbines 
manufactured  by  Brown,  Boveri  &  Company,  Baden,  Switzer- 
land. The  arrangement  of  the  turbine  blades  and  the  balancing 


Fig.     3.     Governor  of  Brown-Boveri  Turbine. 

pistons,  including  the  increased  diameter  of  the  drum  at  the  low- 
pressure  end,  are  substantially  the  same  as  in  the  Parsons  turbine 
as  made  by  the  Westinghouse  Machine  Company  in  this  country. 
The  governor,  which  is  shown  in  Fig.  13,  also  operates  on  the 
same  principle,  following  the  lines  laid  down  in  the  Parsons 
patent  in  1896.  The  oiling  is  by  forced  lubrication  and  the  gen- 
eral arrangement  of  the  valves,  governor,  oil  piping,  etc.,  is  clearly 
indicated  in  the  engraving. 

By-pass  Valves. — Mr.  Brown  of  this  firm  has  taken  out  a 
United  States  patent  for  a  by-pass  valve  to  supply  high-pressure 
steam  to  the  low-pressure  end  of  the  turbine,  which  is  somewhat 


152 


STEAM  TURBINES 


different  from  the  Hodgkinson  valve  already  illustrated.  The 
invention  consists  in  providing  pipes  from  the  steam  chest  to  the 
intermediate  stages  of  the  turbine  and  controlling  the  opening  of 
these  pipes  by  the  governor. 

In  Fig.  15,  steam  is  admitted  through  the  supply  pipe,  A,  and 
when  the  throttle  valve,  B,  is  raised  by  the  governor,  it  will  pass 
in  the  direction  of  the  arrows  through  the  valve,  C,  to  the  steam 
space,  D ,  whence  it  will  enter  the  high-pressure  end  of  the  turbine. 


Fig.    15.     By-Pass  Arrangement. 

The  pipes,  E,  F  and  G,  leading  to  other  points  in  the  turbine,  are 
opened  to  the  steam  space  by  movement  of  the  valve,  C,  which 
uncovers  several  ports  shown  in  succession,  when  it  is  raised. 
Ordinarily,  it  is  kept  in  its  lowest  position  by  the  spring,  S,  thus 
closing  all  the  ports.  If,  however,  the  speed  of  the  turbine 
decreases  under  a  heavy  load  the  governor  will  admit  a  greater 
quantity  of  steam  through  valve  D  to  the  stearn  space,  increasing 
the  pressure  in  the  steam  space  and  forcing  the  piston,  P,  upward 
against  the  pressure  of  the  spring,  by  this  means  raising  the  valve, 
C,  and  admitting  steam  successively  to  the  different  stages  of  the 
turbine.  Brown's  patent  also  covers  an  arrangement,  for  a  similar 
purpose,  having  the  valve,  C,  controlled  directly  by  the  governor. 


REACTION  TURBINES  153 

The  Allis-Chalmers  Turbine. 

The  illustration,  Fig.  16,  shows  the  first  turbine  installed  in  this 
country  of  this  'type.  It  is  now  in  operation  at  Utica,  N.  Y.  It  is 
rated  at  1,500  Kw.,  normal  load,  and  runs  at  a  speed  of  1,800  revo- 
lutions per  minute.  It  is  direct-coupled  to  an  Allis-Chalmers  two- 
phase,  60-cycle  revolving-field  alternator,  operating  at  2,500  volts. 
The  unit  has  a  continuous  overload  capacity  of  25  per  cent,  with 
a  3-hour  50  per  cent  overload  capacity  without  exceeding  a  safe 
generator  temperature,  and  capable  of  a  100  per  cent  safe  mo- 
mentary overload.  Artificial  ventilation  by  means  of  an  elec- 
trically driven  fan  blower  will,  however,  enable  the  unit  to  be  run 
safely  beyond  its  rated  overload  capacity. 

Bidding. — The  chief  distinguishing  feature  is  the  blading.  The 
roots  of  the  blades  are  formed  in  dovetail  shape  by  special  ma- 
chinery, and  are  inserted  in  slots  cut  in  foundation  or  base  rings ; 
these  slots  being  formed  by  special  machine  tools  in  such  a  way  as 
to  exactly  conform  to  the  shapes  of  the  blade  roots.  The  founda- 
tion rings  themselves  are  of  dovetail  shape  in  cross  section  and  are 
inserted  in  dovetailed  grooves  cut  in  the  turbine  casing  and  spin- 
dle respectively,  in  which  they  are  held  by  key  pieces,  much  in  the 
same  way  that  the  well-known  "Lewis  bolt"  is  fastened.  In  order 
to  further  insure  the  integrity  of  the  construction,  the  key  pieces 
or  rings  after  being  driven  into  place  are  upset  into  undercut 
grooves. 

Another  noticeable  feature  of  the  blading  is  the  method  of  rein- 
forcing and  protecting  the  tips  of  the  blades,  which  is  a  point  upon 
which  much  thought  has  been  expended  by  various  inventors.  In 
forming  the  blades  a  shouldered  projection  is  left  at  the  tip.  This 
is  inserted  in  a  slot  punched  in  a  shroud  ring;  the  slots  being 
punched  by  special  machinery  in  such  a  way  as  to  produce  ac- 
curate spacing  and  at  the  same  time  form  the  slots  so  that  they 
will  give  the  proper  angles  to  the  blades  independent  of  the  slots 
in  the  base  ring.  After  the  blade  tips  are  inserted  in  the  slots  in 
the  shroud  rings  they  are  riveted  over  by  specially  arranged  pneu- 
matic machinery. 

The  shroud  rings  are  channel  shaped  with  outwardly  projecting 
flanges  which,  after  assembly  in  the  turbine,  are  turned  and  bored 
to  give  the  necessary  working  clearance.  The  flanges  of  the  chan- 


REACTION  TURBINES 


155 


nels  are  made  so  thin  that,  although  amply  sufficient  for  stiffness, 
the  shroud  ring  does  not  have  the  disadvantage  of  a  solid  shroud 
which  acquires  a  dangerous  temperature  by  friction  in  case  of  an 
accidental  contact  of  the  rotating  and  stationary  parts.  The  use  of  a 
protecting  shroud  ring  not  only  stiffens  the  blades,  but  enables  the 
working  clearance  to  be  made  smaller  than  in  the  case  of  naked 
blade  tips,  without  danger  in  case  of  accidental  contact,  thus  re- 


CYUNDE.R      I 

zri 


Fig.    17.     Scheme   of   Blading. 

ducing  the  leakage  loss  to  a  minimum.  The  shroud  also  acts  as  a 
safety  device,  protecting  the  blades  in  case  of  contact  between  sta- 
tionary and  rotating  members,  and  preventing  any  individual  blade 
from  working  loose  and  causing  damage. 

The  entire  blading  is  produced  by  machinery  and  is  made  up  in 
half  rings  in  the  blading  shop  and  carefully  inspected  before  being 
inserted  in  the  turbine.  Fig.  17  shows  the  general  scheme  of  the 
blading. 

Balance  Pistons. — Another  special  feature  of  this  turbine  is  the 


156 


STEAM  TURBINES 


Fig.  18.     Two  Half-rings  of  Blades — the  Largest  and  the  Smallest  in 
the  Turbine. 

absence  of  the  usual  low-pressure  "balance  piston."  Instead  of 
being  at  the  high-pressure  end,  as  formerly  arranged,  it  is  at  the 
low-pressure  end,  and  by  making  this  piston  in  such  a  way  that  its 
circular  area  is  equal  to  the  annular  area  of  the  pistons  used  in  the 
older  construction,  the  low-pressure  balance  piston  can  be  much 
smaller.  Instead  of  reducing  the  leakage  past  this  piston  by  means 
of  "dummy  packing,"  as  in  the  high-pressure  and  intermediate  pis- 
tons, and  as  used  in  the  low-pressure  pistons  of  the  older  construc- 


Fig.  19.     From  a  Photograph,  Showing  Uniformity  of  Blading. 


158  STEAM  TURBINES 

tion,  a  labyrinth  packing  of  radial  baffling  type  has  been  adopted, 
thus  eliminating  small  axial  clearance  in  this  turbine.  The  advan- 
tage claimed  for  this  construction  is  the  use  of  smaller  working 
clearances  in  the  high-pressure  and-  intermediate  balance  pistons. 
Willans  &  Robinson  Turbine. — The  Allis-Chalmers  Company 
have  effected  an  alliance  with  the  Turbine  Advisory  Syndicate  of 
England,  which  includes  Messrs.  Willans  &  Robinson,  the  high- 
speed engine  builders  of  Rugby,  and  several  other  well-known 
English  firms.  Their  turbines  are  similar  in  construction  to  those 
manufactured  by  Willans  &  Robinson,  both  being  made  under  the 
Fullagar  patents.  In  Fig.  20  is  a  Willans  &  Robinson  turbine 
with  the  top  casing  removed  and  which  shows  certain  of  the 
features  of  construction  more  clearly  than  the  illustrations  pre- 
viously given.  In  this  figure  A  is  the  rotor,  B  the  top  casing,  and 
C  the  admission  valve.  D  is  the  steam  chest  in  which  is  the  valve 
controlled  by  the  governor,  admitting  steam  through  the  pipe  E 
to  the  high-pressure  end  and  through  pipe  F  to  the  low-pressure 
end  in  case  of  overloads.  At  G  are  two  of  the  balance  pistons, 
while  at  H  is  the  small  low-pressure  balance  piston  previously  re- 
ferred to  in  the  description  of  the  Allis-Chalmers  turbine.  It  will 
be  noted  that  all  connections,  such  as  pipes  and  valves,  are  attached 
to  the  lower  half  of  the  casing,  so  that  the  upper  half  may  be  re- 
moved without  disturbing  any  of  the  fittings. 


CHAPTER  VIII 


MISCELLANEOUS  TURBINES  AND  APPARATUS. 


Combined  Impulse  and  Reaction  Turbines. 

Illustrations  have  already  been  given  of  several  types  of  impulse 
and  reaction  turbines  and  we  will  now  consider  two  different  ma- 
chines which  combine  the  impulse  and  reaction  principles.  Ob- 


Fig.   1.     Combined   Impulse   and   Reaction  Turbine. 

jections  to  turbines,  like  the  Parsons,  which  act  on  the  reaction 
principle,  exist  in  the  great  overall  length,  the  mechanical  diffi- 
culties in  the  construction  of  the  rotor,  and  the  necessity  for 
balance  pistons  to  take  up  the  end  thrust.  By  making  the  tur- 
bine double  flow — that  is,  by  admitting  steam  at  the  center  and 
passing  it  through  equal  rows  of  blades  on  each  side — the  end 
thrust  can  be  eliminated,  but  the  overall  length  becomes  much 
greater  and  there  is  more  loss  from  friction  and  leakage. 

British- Westinghouse  Turbine. — In  this  type  of  machine,  which 


160  STEAM  TURBINES 

has  also  been  experimented  with  by  The  Westinghouse  Machine 
Company  in  this  country,  the  double-flow  arrangement  is  adopted 
t@  eliminate  end  thrust  and  a  long  drum  is  obviated  by  having 
the  turbine  divided  into  two  or  more  stages,  in  the  first  of  which 
the  steam  acts  on  the  impulse  principle  and  in  the  second,  or 
last,  of  which  it  acts  on  the  reaction  principle.  Fig.  1,  taken  from 
the  patent  records,  shows  the  general  features  of  the  design, 
although  the  turbine  as  actually  constructed  is  different  in  some 
of  the  details. 

Steam  enters  through  the  valve  at  the  center  and  flows  to  the 
right  and  left  through  nozzles  (not  shown)  in  which  it  is  ex- 
panded to  a  considerably  lower  pressure.  The  nozzles  direct  the 
steam  against  a  series  of  impulse  blades  at  A,  consisting  of  two 
rows  of  moving  and  one  row  of  stationary  vanes.  It  then  passes 
to  the  longer  series  of  blades  which  may  be  entirely  of  the  reac- 
tion or  Parsons  type ;  or  may  start  at  B  with  several  rows  of  im- 
pulse vanes,  divided  into  two  or  more  pressure  stages,  and  end 
with  the  reaction  blading  at  C.  A  feature  of  the  design  is  that 
the  drum  is  of  constant  diameter,  instead  of  in  steps,  as  in  the 
regular  Parsons  turbine,  the  reaction  part  of  the  drum  cor- 
responding to  the  last  stage  or  step  in  the  Parsons  turbine  and 
the  expansion  previous  to  this  point  being  taken  care  of  in  the 
impulse  section  of  the  turbine. 

As  made  by  the  British  company,  the  drum  or  cylinder  is  a 
single  forging  carried  on  the  spindle  by  a  central  supporting  disk 
and  stiffening  disks  of  thin  steel  fitted  into  the  ends.  The  impulse 
blades  are  carried  by  steel  rings  shrunk  onto  the  drum  and  the 
reaction  blades  by  grooves  in  the  drum  itself.  Steam  is  expanded 
to  about  60  pounds  in  the  impulse  section  by  means  of  the  nozzles 
and  two  rows  of  rotating  blades. 

A  vertical  turbine,  combining  both  the  impulse  and  reaction 
principles,  is  manufactured  by  the  Union  Machine  Company, 
Essen,  Germany,  and  a  horizontal  one  is  manufactured  by  Sulzer 
Brothers,  Winterthur,  Switzerland.  A  sectional  elevation  of  the 
latter  is  shown  in  Fig.  2.  Steam  enters  through  valve  A,  passing 
through  the  two  impulse  wheels  at  B.  It  then  flows  through  the 
passages  of  the  reaction  wheel,  C,  thence  around  to  the  other  side 
through  the  passages  of  reaction  wheel  D,  to  the  exhaust  space  E. 


162 


STEAM  TURBINES 


It  is  designed  that  the  thrust  of  reaction  wheel  C  shall  be  balanced 
by  that  of  wheel  D.  The  admission  valve  A  is  an  oscillating  valve 
controlled  by  the  piston  P,  which  in  turn  is  governed  by  the  aux- 
iliary valve  V.  This  latter  valve  is  given  an  oscillating  motion  by 
the  wormwheel  W  and  the  position  of  the  valve  is  at  all  times  de- 
termined by  the  position  of  the  governor. 

The  Lindmark  Steam  Turbine. 

It  has  been  known  for  some  time  that  the  De  Laval  Steam 
Turbine  Company,  Stockholm,  Sweden,  were  developing  a  com- 


Fig.    3.     Lindmark  Turbine. 

pound  steam  turbine  for  .use  where  greater  power  is  desired  than 
it  has  been  found  advisable  to  attempt  to  supply  with  the  single- 
wheel  De  Laval  turbine.  This  new  turbine  is  made  under  the 
patents  of  T.  G.  E.  Lindmark,  the  first  of  which  was  issued  in  this 
country  in  1902.  Along  substantially  the  same  lines  as  the  Lind- 
mark inventions  are  patents  taken  out  by  P.  J.  Hedlund  and  con- 
signed to  the  same  company.  At  present  no  information  con- 
cerning the  Lindmark  turbine  is  available  other  than  given  in  the 
patent  records,  but  inasmuch  as  a  new  principle  is  involved  in  the 
operation  of  this  invention,  it  will  be  of  interest  to  explain  the 
features  of  this  turbine  in  so  far  as  is  now  possible. 

Fig.  3  is  a  representation  of  a  typical  turbine  containing  the 
turbine  wheels,  A,  B,  C,  D  and  E,  attached  to  the  shaft,  S,  and 


MISCELLANEOUS  TURBINES  AND  APPARATUS 


163 


rotating  with  it.  Each  wheel  rotates  in  a  separate  compartment 
and  is  of  the  radial  outflow  type.  The  steam  enters  through  an 
inlet  pipe,  /,  to  the  steam  space,  G,  whence  it  flows  in  the  direction 
of  the  arrow  into  the  first  wheel,  A.  The  two  sides  of  the  wheel 
converge  at  the  periphery,  forming  a  contracted  outlet,  and  the 
steam  discharges  between  curved  vanes  located  in  this  contracted 
space,  and  flows  through  a  diverging  annular  opening  shown  at 
HH,  which  guides  the  steam  into  a  second  compartment  at  KK. 
This  process  is  repeated  in  connection  with  wheels,  B,  C,  D  and 
E. 


Fig.  4.     Showing  Principle  of  Nozzle  Action. 

Purpose  of  the  Turbine. — The  peculiarity  of  the  turbine  lies  in 
the  pressure  and  velocity  effect  upon  the  steam  of  the  converging 
and  diverging  passages  as  the  steam  progresses  through  the  tur- 
bine. The  purpose  of  the  inventor  is  to  transform  the  potential 
energy  and  pressure  of  the  steam  into  kinetic  energy  and  velocity 
as  it  flows  through  the  converging  passages  of  any  given  wheel, 
and  then  to  transform  the  residual  kinetic  energy  and  velocity  of 
the  steam  after  it  discharges  from  the  wheel  into  potential  energy 
and  pressure  during  the  passage  through  the  annular  space,  HH. 

Flow  of  Steam  Through  Nozzles. — In  order  to  explain  the  prin- 
ciple whereby  this  result  is  accomplished,  it  will  be  clearer  to 
first  refer  to  the  action  of  steam  in  flowing  through  nozzles  of  dif- 
ferent proportions.  In  Fig.  4  is  shown  a  converging  and  di- 
verging nozzle  at  A,  through  which  steam  flows  from  a  higher  to 


164  STEAM  TURBINES 

a  lower  pressure  in  the  direction  of  the  arrow.  The  nozzle  is 
supposed  to  be  proportioned  so  as  to  give  complete  expansion,  and 
the  pressure  will  accordingly  drop  from  a  to  b,  while  the  velocity 
will  increase  from  c  to  d  as  indicated  by  the  diagram  below  the 
nozzle. 

At  B  the  diverging  part  of  the  nozzle  is  cut  away  so  that  the 
steam  does  not  have  the  opportunity  to  fully  expand  until  it  has 
left  the  nozzle.  The  pressure,  therefore,  will  gradually  drop  from 
a  to  b,  at  which  latter  point  the  steam  leaves  the  nozzle  and  sud- 
denly expands  to  the  lower  external  pressure,  producing  a  sudden 
change,  b,  in  the  curvature  of  the  pressure  line.  Inasmuch  as  the 
expansion  is  not  complete  in  the  nozzle,  the  velocity  line  does  not 
rise  to  as  high  a  point  as  in  the  previous  case. 

At  C  the  nozzle  is  shown  lengthened  out  so  as  to  produce  over- 
expansion;  that  is  to  say,  when  the  steam  reaches  point  x  in  the 
nozzle  it  is  expanded  down  to  the  external  pressure  of  the  medium 
into  which  the  nozzle  discharges,  and  beyond  this  point  the  steam 
will  expand  a  few  pounds  below  the  outside  pressure  until,  say, 
point  y  is  reached.  After  this  the  pressure  will  rapidly  rise  again 
until  it  reaches  the  external  pressure  at  the  mouth  of  the  nozzle. 

The  Action  in  the  Lindmark  Turbine. — It  is  this  latter  action 
which  takes  place  in  the  diverging  passages  of  the  Lindmark 
turbine. 

The  application  of  the  principle  can  best  be  explained  by  re- 
ferring to  the  detail  sectional  sketch  of  the  Lindmark  turbine 
shown  in  Fig.  5,  which  represents  half  of  one  of  the  wheels  and 
the  connecting  passages.  Steam  enters  through  the  inlet  and 
passes  through  the  wheel  in  the  direction  of  the  arrow,  and  thence 
to  the  exhaust  chamber,  whence  it  escapes  through  the  valve  B. 
Suppose  first  the  turbine  wheel  to  be  blocked  so  that  it  will  not 
turn,  valves  A  and  B  to  be  wide  open,  and  the  steam  to  flow  freely 
through  the  turbine.  At  points  1,  2,  3,  4  and  5,  are  openings  to 
which  gauges  may  be  attached  for  determining  the  pressure  at  the 
different  points.  The  steam  passages,  C  and  D,  will  now  act  like  a 
converging  and  diverging  nozzle.  At  1  the  entering  steam  will 
be  at  boiler  pressure;  at  2  the  pressure  within  the  wheel  casing 
will  be  substantially  that  at  the  throat  of  the  nozzle,  or  about 
%o  of  the  boiler  pressure;  at  3  it  will  be  slightly  lower;  and 


MISCELLANEOUS  TURBINES  AND  APPARATUS 


165 


at  4  nearly  as  low  as  the  pressure  in  the  exhaust  space  indicated  by 
the  gauge  at  5.     (See  "Steam  Nozzles,"  Chap.  I.) 

Action,  Wheel  Blocked,  Exhaust  Valve  Partially  Closed. — 
Now,  suppose  valve  B  to  be  closed  as  much  as  possible  without 
increasing  the  pressure,  2,  within  the  wheel  chamber.  The  pres- 
sure at  1  will  obviously  remain  as  before,  and  the  expansion  in  the 
passages,  C,  of  the  wheel  will,  as  before,  carry  the  pressure  at  the 
throat  of  the  nozzle  to  %0  of  the  initial  pressure.  After  the 
steam  enters  the  diverging  space,  D,  however,  over-expansion  will 
occur  just  as  in  nozzle  C,  Fig.  4.  Probably  the  lowest  pressure 


EXHAUST 


Fig.  5.     Enlarged  View  of  Passages. 

will  be  at  or  near  the  point  where  the  gauge  is  attached  at  3,  and 
from  there  on  the  velocity  will  decrease  and  the  pressure  will  in- 
crease so  that  the  gauges  at  4  and  5  will  indicate  a  pressure  con- 
siderably higher  than  that  in  the  wheel  chamber  at  2,  but  lower 
than  the  initial  pressure  at  1. 

Action  with  the  Wheel  Rotating. — Finally  if  the  wheel  now 
be  supposed  to  turn,  instead  of  being  blocked  as  before,  and  the 
exhaust  valve  still  be  partially  closed,  the  principle  of  the  steam's 
action  will  remain  the  same,  the  only  change  being  due  to  the 
altered  velocity  of  the  steam  owing  to  the  fact  that  part  of  the 
velocity  will  be  absorbed  by  the  vanes  of  the  rotating  wheel. 


166  STEAM  TURBINES 

Under  these  conditions  the  velocity  of  flow  when  the  steam  reaches 
the  entrance  to  the  diverging  portion  will  be  less  than  before  and 
this  will  produce  an  effect,  which,  according  to  the  experiments 
of  Lindmark,  will  reduce  the  pressures  somewhat  at  points  4  and  5. 
The  Lindmark  turbine  is"  a  reaction  turbine,  since  the  expansion 
of  the  steam  and  increase  in  velocity  occur  in  the  passages  of  the 
wheel  vanes  and  the  curvature  given  to  them  is  similar  to  that 
employed  for  reaction  wheels  of  other  types. 

THE  RATEAU  STEAM  ACCUMULATOR  SYSTEM. 

Turbines  for  Low-Pressure  Steam. — One  of  the  most  interest- 
ing applications  of  the  Rateau  turbine  is  in  employment  with  low- 
pressure  steam  supplied  from  engines  working  intermittently,  such 
as  rolling-mill  engines,  hoisting  engines,  etc.  The  majority  of 
such  engines  are  working  under  wasteful  conditions.*  They 
usually  operate  under  widely  varying  loads,  frequently  with  only 
a  small  degree  of  expansion  and  usually  exhausting  freely  into 
the  air.  Professor  Rateau  has  given  special  attention  to  the  em- 
ployment of  waste  steam  from  such  engines  and  has  obtained 
satisfactory  results  by  means  of  his  regenerative  accumulator  of 
steam  combined  with  low-pressure  turbines.  The  accumulator  is 
intended  to  regulate  the  intermittent  flow  of  steam  before  it  passes 
to  the  turbine,  and  consists  essentially  of  a  tank  containing  solid  or 
fluid  materials  which  play  the  part  of  a  flywheel  for  heat 

Description  of  the  Accumulator. — In  his  paper  before  the 
American  Society  of  Mechanical  Engineers  in  June,  1904,  Pro- 
fessor Rateau  gave  the  following  description  of  his  accumulator: 

The  steam  collects  and  is  condensed  as  it  arrives  in  large  quantities  in  the 
apparatus,  and  is  again  vaporized  during  the  time  when  the  exhaust  of  the 
principal  engine  diminishes  or  ceases.  The  necessary  variations  for  con- 


*Trials  of  modern  winding  engines  frequently  show  a  steam  consumption  of  65  to  80 
pounds  per  horse-power  in  mineral  hoisted,  while  100  to  120  pounds  is  not  uncommon, 
so  that  allowing  20  per  cent  condensation  in  the  cylinders  and  passages  of  the  engine, 
there  is  discharged  into  the  atmosphere  by  such  an  engine  a  minimum  of  9,000  to 
10,000  pounds  of  steam  per  hour  which  is  totally  lost.  This  steam  is  theoretically 
capable  of  developing  500  to  600  horse-power  if  supplied  at  atmospheric  pressure  and 
exhausted  at  a  vacuum  of  27  inches.  Rolling  mill  engines  frequently  consume  50,000 
pounds  of  steam  per  hour,  which  is  theoretically  capable  of  developing  over  2,500  horse- 
power in  expanding  from  atmospheric  pressure  to  the  vacuum  of  an  ordinary  condenser, 
even  after  deducting  20  per  cent  for  condensation  losses. — From  paper  by  Leonce 
Battu,  read  at  a  meeting  of  the  Western  Society  of  Engineers,  September,  1904. 


MISCELLANEOUS  TURBINES  AND  APPARATUS 


167 


densation  and  regeneration  of  the  steam  correspond  to  fluctuations  in  pres- 
sure in  the  accumulator,  this  pressure  rising  when  the  apparatus  is  being 
filled  and  descending  when  it  is  discharging  into  the  turbine.  Water  which 
has  a  very  high  calorific  capacity  has  been  used  as  a  heat  fly-wheel,  but  in 
order  to  rapidly  communicate  to  a  liquid  mass  a  considerable  quantity  of 


Exhaust-Steam  from 
the  Winding-Engine 


Waste-water 


FLAN  OF  ONE-THIRD  OF  A  BASON 


Fig.  6.     Rateau  Steam  Accumulator  with  Iron  Trays. 


heat  corresponding  to  the  latent  heat  of  steam  to  be  condensed  it  becomes 
necessary,  owing  to  the  poor  conductivity  of  water,  either  to  arrange  it  in 
thin  layers  or  to  cause  a  rapid  circulation  in  order  to  increase  the  surface 
of  contact  between  the  steam  and  the  water  itself.  The  first  solution  of  the 
problem  gave  rise  to  the  accumulator  with  flat  cast-iron  trays  in  which  wa- 
ter is  contained  in  shallow  vessels  arranged  one  above  the  other,  Fig.  6. 
The  second  solution  of  the  problem  gave  rise  to  the  accumulator  with  wa- 
ter only  in  which  a  rapid  circulation  was  produced  by  the  injection  of  steam 


168 


STEAM  TURBINES 


into  the  body  of  the  liquid  itself,  Fig.  7.  The  low-pressure  turbine,  fed  by 
the  regular  flow  which  comes  from  the  accumulator,  and  working,  for  ex- 
ample, between  an  admission  pressure  of  15  pounds  per  square  inch  and  a 
vacuum  at  the  condenser  of  27  inches  of  mercury  (back  pressure  of  1.6 
pounds)  can  furnish  an  electric  horse-power  for  about  31  pounds  of  steam 
per  hour.  In  steel  works,  where  "reversible  steam  rolls  are  employed  con- 
suming about  45,000  pounds  of  steam  per  hour,  it  will  be  easy  to  develop, 
by  means  of  accumulators  and  turbines,  an  extra  output  of  over  1,100  elec- 
tric horse-power. 

Accumulators   are   fitted   with   several   accessories   which   are 
necessary   for  their  successful   operation.     One  of  these  is   an 


SECTION  edefgh  SECTION  ab 

Fig.    7.     Accumulator    with    Water    only. 

automatic  relief  valve  to  allow  the  steam  from  the  engine  to 
escape  into  the  atmosphere  of  the  condenser  if  the  turbine  should 
not  require  all  of  the  steam  exhausted  by  the  main  engine.  An- 
other is  an  automatic  expansion  valve  provided  so  that  live  steam 
from  the  boiler  may  be  admitted  to  the  turbine,  should  the  main 
engine  be  temporarily  shut  down  or  if  sufficient  exhaust  steam  is 
not  available.  There  should  also  be  a  steam  check  valve  and  a 
water  check  valve,  the  former  for  shutting  off  the  accumulator 
from  the  turbine  when  the  main  engine  is  shut  down  and  the  tur- 
bine is  supplied  with  live  steam  only;  while  the  latter  is  used  to 
prevent  the  water  in  the  accumulator  returning  toward  the  main 


MISCELLANEOUS  TURBINES  AND  APPARATUS          169 

engine,  through  the  exhaust  supply  pipe,  when  the  engine  is  shut 
down. 

Calculations  for  an  Accumulator* — To  calculate  the  weight  of 
water  or  cast  iron  necessary  for  an  accumulator,  we  must  know : 

( 1 )  The  total  weight  of  steam  required  by  the  turbine  per  hour ; 

(2)  assume  a  length  of  time  for  the  accumulator  to  operate  with- 
out a  steam  supply;  and  (3)  assume  an  allowable  dr,op  in  tem- 
perature of  the  accumulator  while  the  steam  supply  is  shut  off  and 
the  turbine  is  in  operation.     For  example,  suppose  the  turbine  to 
consume  2,200  pounds  of  steam  per  hour;  the  duration  of  the  stop 
to  be  one  minute ;  and  the  range  of  temperature  10  degrees  F. 

Let  G=weight  of  steam  used  by  turbine  during  stoppage  of 
steam  supply. 

2,200X1 

ThenG=-         -=36.6. 
60 

At  atmospheric  pressure  one  pound  of  steam  contains  966 
B.  T.  U.  Whence,  the  accumulator  must  be  able  to  deliver 

36.6X966=35,400  B.  T.  U. 

Since  the  specific  heat  of  water  is  one,  the  weight  of  water 
necessary  to  deliver  this  quantity  of  heat  with  a  range  of  tempera- 
ture of  10  degrees  F.  is 

35,400 
=3,540  pounds,  or  about  1.5  tons. 

If  the  accumulator  were  to  contain  cast  iron  instead  of  water, 
the  specific  heat  of  which  is  0.11,  a  weight  of  iron  would  be  re- 
quired about  9  times  as  great  as  for  water.  Hence,  for  cast  iron 
we  have 

1.5 

—=13.75  tons. 
0.11 

Tests  on  Rateau's  Accumulator  System. — The  first  plant  of  this 
type  to  be  installed  was  at  the  Bruay  Mines,  France.  The  ac- 
cumulator was  of  the  type  shown  in  Fig.  6,  and  a  drawing  of  the 

"Paper  by  Leonce  Battu,  Western  Society  of  Engineers,   Sertember,   1904. 


MISCELLANEOUS  TURBINES  AND  APPARATUS          171 

complete  plant  is  shown  in  Fig.  8.  This  plant  is  of  3,000  horse- 
power, and  tests  taken  at  the  time  the  machinery  was  installed  and 
15  months  afterwards  showed  no  diminution  of  efficiency.  In  one 
series  of  tests  the  pressures  of  the  steam  supplied  to  the  turbine 
ranged  from  12  to  14.5  pounds  absolute,  and  the  condenser  pres- 
sures ranged  from  2.13  to  2.62  pounds  absolute.  The  consumption 
of  steam  per  horse-power  hour,  the  horse-power  being  the  electric 
horse-power  delivered  by  the  generator,  ranged  from  37.4  to  45.2. 
In  his  paper  previously  referred  to  M.  Battu  gives  the  following 
estimate  of  the  probable  consumption  of  steam  operating  on  this 
system : 

The  table  shows  the  steam  consumption  required  to  give  a  horse- 
power electric  at  the  terminals  of  the  generators: 


30  Ibs. 

15  Ibs. 

7.5  Ibs. 

Vacuum  in  the  Condenser  of  27  6  inches                 111  Ibs 

21.7 

26.4 

8J.8 

"          "    "             "             "   sW  0      "              .   ...1.85    u 

23.6 

31.7 

«7.3 

"          "     "             "             "   21  6      "                      2  56    " 

26.4 

36.3 

61.6 

The  first  horizontal  line  of  the  table  corresponds  to  the  case  where  a  sur- 
face condenser  giving  a  vacuum  of  27.6  inches  is  used.  The  second  to  an 
ordinary  jet  condenser  giving  a  vacuum  of  26  inches,  and  the  third  to  an 
ejector  condenser  giving  a  vacuum  of  24.6  inches.  These  vacuums  are 
easily  obtained  in  practice. 

The  central  vertical  column  corresponds  to  the  usual  conditions  when 
the  steam  is  supplied  to  the  turbine  at  atmospheric  pressure.  The  first 
column  gives  the  consumption  in  the  case  where  the  primary  engines  ex- 
haust against  a  gauge  pressure  of  fifteen  pounds,  in  order  to  take  advantage 
of  the  superior  efficiency  with  which  it  can  be  used  in  the  turbine,  while  the 
third  column  represents  the  condition  where  it  is  desirable  not  to  disturb 
the  condensing  operation  of  the  main  engines,  and  consequently  to  utilize 
steam  exhausting  from  them  at  about  7^2  pounds  absolute. 

Low-Pressure  Curtis  Turbine. 

The  Philadelphia  Rapid  Transit  Company  have  installed  an  800 
Kw.  Curtis  low-pressure  turbine  in  their  station  at  Thirteenth  and 
Mt.  Veriaon  Streets.  This  station  is  equipped  with  Corliss  en- 
gines, and  as  it  is  located  midway  between  the  Schuylkill  and 
Delaware  rivers  the  engines  have  always  been  run  non-condensing. 
An  Alberger  condenser  and  cooling  tower  have  been  installed, 


172  STEAM  TURBINES 

however,  for  use  with  the  turbine  and  a  vacuum  of  28  inches  or 
more  has  been  maintained  during  the  cool  weather.  The  turbine 
has  four  wheels,  each  with  a  single  row  of  buckets.  When  the 
turbine  is  receiving  steam  at  atmospheric  pressure,  without  moist- 
ure, the  guarantees  provide  that  the  steam  consumption  shall  not 
exceed  36  pounds  per  kilowatt  hour  at  full  load  and  40  pounds  at 
half  load,  back  pressure  being  two  inches.  At  four  inches  back 
pressure  these  figures  are  respectively  45  and  50  pounds.  Tests 
made  on  the  machine  at  the  factory  showed  even  better  per- 
formance. It  is  estimated  that  the  turbine  will  increase  the  out- 
put of  the  engine  or  engines  that  supply  steam  to  it  about  66% 
per  cent  instead  of  the  25  per  cent  usually  expected  from  the 
application  of  a  condenser. 


CHAPTER  IX 

STEAM    TURBINE    PERFORMANCE— COMPARISONS    WITH 
THE  STEAM  ENGINE. 

There  will  be  found  in  this  chapter  the  results  of  a  number  of 
tests  upon  turbines  of  different  types.  In  these  the  steam  con- 
sumption is  usually  given  in  pounds  per  electrical  horse-power 
per  hour,  or  in  pounds  per  kilowatt  hour.  In  the  majority  of 
cases  turbines  are  direct-connected  to  electric  generators,  and  the 
power  is  most  readily  measured  by  the  electrical  instruments  of 
the  switchboard,  which  show  the  output  of  the  generator  instead  of 
the  actual  power  developed  by  the  turbine.  In  factory  tests,  how- 
ever, before  the  turbine  is  shipped,  the  brake  horse-power  is  often 
determined,  since  means  are  usually  at  hand  for  attaching  an 
absorption  dynamometer.  This  plan  is  followed  at  the  works  of 
the  Westinghouse  Machine  Company,  Pittsburg,  Pa. 

Kilowatts  and  Electrical  Horse-Power. — The  power  delivered 
by  the  generator  is  expressed  in  kilowatts  or  in  electrical  horse- 
power, the  latter  being  the  equivalent,  in  electric  units,  of  me- 
chanical horse-power.  For  converting  kilowatts  to  horse-power 
and  horse-power  to  kilowatts,  we  have : — 

1  kilowatt=1.3405  horse-power— 1.34  horse-power,  nearly. 

1  horse-power=0.7459  kilowatt=  0.746  kilowatt,  nearly. 

Table  I.  will  be  of  assistance  in  converting  the  more  usual 
values  of  kilowatts  to  electrical  horse-power  and  of  electrical 
horse-power  to  kilowatts. 

Table  II.  gives  steam  consumption  in  pounds  per  electrical 
horse-power  hour  corresponding  to  steam  consumption  per  kilo- 
watt hour,  taken  at  half-pound  intervals,  within  the  limits  usually 
met  with  in  practice. 

Table  III.  gives  steam  consumption  per  kilowatt  hour  corre- 
sponding to  consumption  per  electrical  horse-power  hour. 

Indicated,  or  Internal  Horse-Power. — There  is  no  such  thing 
as  the  indicated  horse-power  of  a  turbine,  because  no  indicator 
has  been,  and  probably  none  can  be,  devised  to  show  the  internal 
power  developed.  While  an  indicator  might  show  the  energy  of  a 


174 


STEAM  TURBINES 


jet  of  steam  discharged  from  a  nozzle,  it  would  be  practically 
impossible  to  register  the  amount  of  energy  given  up  by  a  jet  to 
the  blades  of  a  compound  turbine,  where  the  losses  might  be 
greater  or  less,  according  to  the  design,  load,  and  other  running 
conditions. 

Engineers  are  so  familiar  with  the  water  rates  of  reciprocating 
engines  on  the  basis  of  the  indicated  horse-power,  that  in  com- 
paring a  turbine  with  an  engine  it  is  usual  to  reduce  the  figures 
for  the  steam  consumption  of  the  turbine  to  terms  of  the  indi- 
cated horse-power  of  a  reciprocating  engine  having  the  same 
electrical  output,  or  brake  horse-power,  as  the  case  may  be. 

TABLE  I. 

CONVERSION  OF  HORSE  POWER  INTO  KILOWATTS  AND  KILOWATTS 
INTO  HORSE  POWER. 

1  Kw.  —  1.3405  H.  P.  (1.34). 
1  H.  P.       0.7459  Kw.  (0.746). 


Kilowatts 

Horse 

Kilowatts 

Horse 

Number. 

to  Horse 

Power  to 

Number. 

to  Horse 

Power  to 

Power. 

Kilowatts. 

Power. 

Kilowatts. 

1 

1.34 

0.75 

36 

48.26 

2J.85 

2 

2.68 

1.49 

37 

49.60 

27.60 

3 

4.02 

2.24 

38 

50.94 

28.34 

4 

5.36 

2.98 

39 

52.28 

29.09 

5 

6.70 

3.73 

40 

53.62 

29.84 

6 

8.04 

4.48 

41 

54.96 

30.58 

"J 

9.38 

5.22 

42 

56.30 

31.33 

8 

10.72 

5.97 

43 

57.64 

32.07 

9 

12.06 

6.71 

44 

58.98 

32.82 

10 

13.40 

7.46 

45 

60.32 

33.57 

11 

14.75 

8.20 

46 

61.66 

34.31 

12 

16.09 

8.95 

47 

63.00 

35.06 

13 

17.43 

9.70 

48 

64.34 

35.80 

14 

18.77 

10.44 

49 

65.68 

36.55 

15 

20.11 

11.19 

50 

67.03 

37.30 

16 

21.45 

11.93 

75 

100.5 

55.94 

17 

24.79 

12.68 

100 

134.1 

74.59 

18 

24.13 

13.43 

150 

201.2 

111.9 

19 

23.47 

14.17 

200 

268.1 

149.2 

20 

26.81 

14.92 

250 

335.1 

186.5 

21 

28.15 

15.66 

8)0 

402.2 

223.8 

22 

29.49 

16.41 

350 

469.2 

261.1 

23 

30.83 

17.16 

400 

536.2 

298.4 

24 

33.17 

17.90 

45) 

603.2 

335.7 

25 

33.51 

IS.  65 

500 

6T0.3 

373.0 

26 

34.85 

19  39 

750 

1005. 

559.4 

27 

36.19 

20.14 

1000 

1341. 

745.9 

28 

37.53 

20.88 

1250 

1675. 

932.4 

29 

38.87 

21.63 

1500 

2011. 

1119. 

30 

40.22 

22.38 

2000 

2681. 

1492. 

31 

41.56 

23.12 

2500 

3351. 

1865. 

32 

42.90 

23.87 

3000 

4022. 

223S. 

33 

44.24 

24.61 

4000 

5362. 

2984. 

34 

45.58 

25.36 

5000 

(703. 

3730. 

35 

46.92 

26.11 

STEAM  TURBINE  PERFORMANCE 
TABLE  II. 

RELATIVE  STEAM  CONSUMPTION  PER  KILOWATT  HOUR  AND 
ELECTRICAL  HORSE  POWER  HOUR. 


175 


Lb.  per 
Kw.  Hour. 

Corresponding 
Ib.  per 
E.  H.  P.  Hour. 

Lb.  per 
Kw.  Hour. 

Corresponding 
Ib  per 
E.  H.  P"  Hour. 

13 

9.70 

24.5 

18  28 

]3.5 

10.07 

25 

18.65 

14 

10.44 

25.5 

19.02 

14.5 

10.82 

26 

19.39 

15 

11.19 

26.5 

19.77 

15.5 

-  11.56 

27 

20.14 

16 

11.93 

27.5 

20.51 

16.5 

12.32 

28 

20.88 

17 

12.68 

28.5 

21.26 

17.5 

13.06 

29 

21.63 

18 

13.43 

29.5 

2^.01 

18.5 

13.80 

30 

22.38 

19 

14.17 

30.5 

22.75 

19.5 

14.55 

31 

23.12 

20 

14.92 

31.5 

23.50 

20.5 

15.29 

32 

23.87 

21 

15.66 

32.5 

24.24 

21.5 

16.04 

33 

24.61 

22 

16.41 

33.5 

24.99 

22.5 

16.79 

34 

25.36 

23 

17.16 

34.5 

25.74 

23.5 

17.53 

35 

26.11 

24 

17.  DO 

TABLE  III. 

RELATIVE  STEAM  CONSUMPTION  PER  ELECTRICAL  HORSE  POWER 
HOUR  AND  KILOWATT  HOUR. 


Lb.  per 
E.  H.  P.  Hour. 

Corresponding 
Ib.  per 
Kw.  Hour. 

Lb.  per 
E.  H.  P.  Hour. 

Corresponding 
Ib.  per 
Kw.  Hour. 

10 

13.40 

19 

25.47 

10.5 

14.08 

19.5 

26.14 

11 

14.75 

20 

26.81 

11.5 

15.42 

20.5 

27.48 

12 

16.09 

21 

28.15 

12.5 

16.76 

21.5 

28.82 

13 

17.43 

22 

29.49 

13.5 

18.10 

22.5 

30.16 

14 

18.77 

23 

30.83 

14.5 

19.44 

23.5 

30.16 

15 

20.11 

24 

3-.M7 

15.5 

20.78 

24.5 

32.84 

16 

21.45 

25 

33.51 

16.5 

22.12 

25.5 

34.18 

17 

22.79 

26 

34.85 

17.5 

23.46 

26.5 

35.52 

18 

24.13 

27 

36.19 

18.5 

24.80 

176  STEAM  TURBINES 

Comparing  Turbine  Performance  with  Engine  Performance* 
— Calculations  of  this  character  must  take  into  account  the  effi- 
ciencies of  engines  and  generators,  data  upon  which  will  shortly 
be  given.  The  actual  calculations  involve  nothing  more  difficult 
than  the  principles  of  percentage. 

Example:  Let  a  turbine  unit  deliver  500  electrical  horse- 
power, and  consume  7,500  pounds  of  steam  per  hour.  Its  rate  of 
steam  consumption  will  then  be  7,500-^500=15  pounds  per  elec- 
trical horse-power  per  hour.  What  would  be  the  indicated  horse- 
power, and  the  consumption  per  indicated  horse-power  per  hour, 
of  a  reciprocating  engine  having  the  same  rate  of  consumption 
per  electrical  horse-power  per  hour?  Assume  the  engine  to  be 
direct-connected  to  a  generator,  the  efficiency  of  the  generator  to 
be  95  per  cent,  and  the  mechanical  efficiency  of  the  engine  94  per 
cent  The  combined  efficiency  will  then  be  0.94X0.95=0.89. 
The  indicated  horse-power=  500-^0. 89=561. 8.  The  steam  con- 
sumption per  indicated  horse-power  per  hour=  7, 500-^561.8= 
13.4  pounds.  The  latter  could  have  been  obtained  directly  by 
multiplying  the  water  rate  for  the  turbine,  15  pounds  per  electrical 
horse-power  per  hour,  by  0.89,  thus :  15X0.89=13.4. 

It  is  to  be  noted  that  in  these  comparative  calculations,  where 
we  estimate  engine  performance  for  comparison  with  turbine  re-- 
suits,  we  use  the  efficiency  of  the  engine-driven  generators,  not  of 
turbine  generators. 


*In  a  paper  upon  the  Curtis  turbine,  read  by  Chas.  B.  Burleigh  before  the  New 
England  Railroad  Club,  April,  1905,  are  calculations  of  the  losses  in  engines  and 
generators,  as  follows:  Let  us  take  a  Curtis  turbine  guarantee  of  20  pounds  of 
steam  per  kilowatt  hour  and  figure  what  the  reciprocating  engine  guarantee  per 
indicated  horse-power  should  be  to  just  equal  it.  Twenty  pounds  per  kilowatt  hour 
is  equivalent  to  20 X -746=  14. 92  pounds  per  electrical  horse-power  hour.  To  start 
with,  we  must  make  cur  turbine  test  with  instruments  mounted  on  the  switchboard, 
and  the  loss  in  the  conductors  from  the  generator  to  the  switchboard  being  1  per 
cent,  we  must  deduct  this  1  per  cent  of  14.92,  or  0.1492  pounds.  14.92 — 0.1492s= 
14.77  pounds.  Next  we  must  deduct  the  generator  loss,  which,  since  the  generator  is 
designed  to  meet  the  engine  speed,  is  in  most  cases  more  than  as  though  the  ideal 
generator  could  have  been  used  and  the  generator  adapted  to  it.  We  should,  therefore, 
allow  at  least  5  per  cent  generator  loss.  Thus,  5  per  cent  of  14.77  =  0.74  and  14.77 — 
0.74=14.03  pounds  per  brake  horse-power.  We  are  now  back  to  the  engine,  but  the 
indicator  card  does  not  take  into  account  the  friction  losses  in  the  engine,  so  these 
must  be  deducted,  and  I  think  you  will  agree  with  me  that  7  per  cent  is  fair  for  this. 
Seven  per  cent  of  14.03=0.98  and  14.03 — 0.98  =  13.05  pounds  per  indicated  horse- 
power. Therefore: 

A  turbine  guarantee  of  20  pounds  of  steam  per  kilowatt  hour, 
An  engine  guarantee  of  13.05  pounds  of  steam  per  indicated  horse-power  hour, 
Or  a  turbine  guarantee  of  14.92  pounds  of  steam  per  electrical  horse-power  hour, 
are  identical. 


STEAM  TURBINE  PERFORMANCE 


177 


Efficiencies  of  Engine-type  Generators. — Table  IV.  has  been 
prepared  from  data  furnished  by  manufacturers  of  generators. 

TABLE  IV. 

EFFICIENCIES  OF  ALTERNATING  AND  DIRECT  CURRENT 
ENGINE-TYPE  GENERATORS. 

ALTERNATING  CURRENT  GENERATORS  (ABOUT  2:300  VOLTS). 


Per  Cent  Efficiency. 

Yt  Load. 

K  Load. 

K  Load. 

Full  Load. 

1#  Load. 

500 

120 

86 

91 

93 

94 

95 

1500 

100 

87 

92 

94 

95 

96 

3000 

75 

68 

93 

95 

96 

96.5 

5000 

75 

89 

94 

95.5 

96.5 

96.7 

DIRECT  CURRENT  GENERATORS. 


Per  Cent  Efficiency. 

K  Load. 

K  Load. 

Yi  Load. 

Full  Load. 

100 

125 

86 

91 

91.5 

92 

500 

250 

88 

92 

93.5 

94 

1000 

250 

89 

92.5 

93.5 

94 

1500 

650 

91 

94 

94.5 

95 

2700 

575 

91 

94 

94.5 

95 

Efficiencies  of  Turbine  Generators. — In  estimating  the  brake 
horse-power  of  a  turbine,  having  given  the  electrical  horse-power, 
or  vice  versa,  there  must  be  an  allowance  for  the  efficiency  of  the 
generator  driven  by  the  turbine. 

In  the  De  Laval  turbine  outfits,  twin  generators  are  used, 
which  reduces  the  size  of  each  generator  by  about  one  half,  and 
the  efficiencies  are  low  on  this  account.  Medium  size  generators 
for  these  turbines,  say  of  200  Kw.  capacity,  have  an  efficiency 
ranging  from  88  to  91  per  cent  between  one  half  and  full  load, 
if  for  direct  current.  In  alternating-current  units  of  the  same 
capacity,  the  efficiency  varies  from  86  to  92  per  cent,  between 
one  half  and  full  load. 

A  test  upon  a  1,250  Kw.  A.  C.  generator  for  use  with  a 
Westinghouse-Parsons  turbine,  reported  by  A.  W.  Mattice  in  the 


178  STEAM  TURBINES 

Electrical  World,  February  20,  1904,  showed  efficiencies  of  86  per 
cent  at  quarter  load,  93  per  cent  at  half  load,  and  96  per  cent  at 
full  load.  Also,  a  report  of  a  test  upon  a  400  Kw.  A.  C.  generator 
for  Parsons  turbine,  normal  voltage  of  440,  by  F.  P.  Sheldon  & 
Co.,  Providence,  R.  I.,  contains  the  following  figures : 

Guaranteed  Efficiency.  Measured  Efficiency. 
Full  load                   94.5  per  cent  96.6  per  cent 

Y±      "  93.5          "  95.7 

y2      "  91  "  94.6 

Tests  on  an  Allis-Chalmers  5,500  Kw.  A.  C.  turbo-generator, 
reported  in  The  Engineer,  February  1,  1906,  showed  the  following 
efficiencies : 

YZ  load,  97      per  cent  full  load,  98.3  per  cent 

Y±      "      97.9         "  1^4      "      98.5 

Alternating  current  generators  of  the  type  used  with  the  Curtis 
turbine,  500  Kw.  and  over,  have  efficiencies  ranging  from  96  to 
97.5  per  cent  at  full  load.  These  generators  have  a  high  electrical 
efficiency,  because  the  Curtis  turbine  runs  at  a  favorable  speed  for 
the  generator;  and  a  high  mechanical  efficiency,  owing  to  the 
small  friction  of  the  vertical  shaft. 

Mechanical  Efficiency  of  Steam  Engines. — The  importance  of 
being  able  to  make  a  just  comparison  between  the  performance 
of  steam  turbines  and  steam  engines  makes  it  essential  to  care- 
fully consider  the  subject  of  engine  friction,  so  that  proper  allow- 
ances may  be  used  when  reducing  electrical  horse-power  to 
equivalent  "internal"  horse-power. 

The  friction  loss  of  steam  engines  remains  very  nearly  con- 
stant at  all  ordinary  operating  loads.  Professor  Thurston  has 
treated  the  subject  exhaustively  in  papers  to  be  found  in  the 
Transactions  of  the  American  Society  of  Mechanical  Engineers, 
volumes  VIII.,  IX.,  and  X.,  and  his  conclusions,  as  well  as  those 
of  others  who  commented  on  his  investigations,  were  that  it  is 
substantially  correct  to  consider  the  friction  loss  constant  under 
varying  loads.  He  found  this  loss  to  be  influenced  to  a  much 
greater  extent  by  the  degree  of  lubrication,  change  in  speed,  steam 


STEAM  TURBINE  PERFORMANCE  179 

distribution,  and  design  and  condition  of  the  engine,  than  by  a 
change  in  load.* 

If  the  friction  load  is  constant,  the  efficiency  of  an  engine  will 
evidently  vary  widely  with  change  in  load.  Thus,  if  a  constant 
friction  load  is  10  per  cent  at  full  load,  it  will  be  20  per  cent  at 
half  load  and  the  efficiencies  will  be  90  and  80  per  cent. 

A  few  tests  will  now  be  quoted  to  indicate  what  allowances  may 
be  made  for  the  mechanical  efficiency  of  engines  of  different  types. 

Friction  Tests  of  Steam  Engines. 

1.  A    5,400   horse-power,   three-cylinder,   vertical,    compound    Westing- 
house  engine**  at  the  Waterside  station  of  the  New  York  Edison  Com- 
pany had  a  combined  efficiency  of  engine  and  generator  of  94.5  to  95.2  per 
cent  under  such  variation  in  load  as  the  engine  was  called  upon  to  carry. 
The  friction  load  was  118.6  horse-power,  or  2.2  per  cent  of  the  normal 
power  of  the  engine. 

2.  A    3,500    horse-power,    triple-expansion,    horizontal    enginef  ^  at    the 
Berlin  electricity  works  had  a  mechanical  efficiency  of  from  89.5  to  92.9 
per  cent  under  running  loads.     When  running  empty  the  friction  horse- 
power was  266.3,  or  7.6  per  cent  of  the  nominal  power  of  the  engine. 

3.  A  2,500  horse-power,   cross-compound,  horizontal   Allis   engine^   at 
the  Harvard  Square  power  station,   Cambridge,   Mass.,  had  a  combined 
efficiency  of  engine  and  generator   (direct  current)   of  90  per  cent  at  an 
electrical  load  of  2,200  horse-power. 

4.  An  850  horse-power,  cross-compound,  horizontal  Rice  and   Sargent 
engine§    (Corliss  type)   gave  the  following  results:     At  normal  load  ihe 
combined  efficiency  of  generator  and  engine  was  93  per  cent;   at   1,000 
horse-power,  94  per  cent ;  at  627  and  490  horse-power,  90  per  cent ;  at  340 
horse-power,  83  per  cent.     The  differences  between   indicated  and  elec- 
trical horse-power  at  the  above  loads,  taken  in  their  order,  were:    57,  59, 
58,  63,  51,  and  57.    The  friction  load,  dynamo  running  idle,  was  45  horse- 
power, or  5.3  per  cent  of  the  rated  power  of  the  engine. 


*See  friction  tests  "No.  4,"  which  follow  in  the  main  text.  In  these  tests  the 
frictional  and  electrical  losses  between  the  cylinder  and  switchboard,  equal  to  the 
differences  between  the  indicated  and  electrical  power  at  different  loads,  vary  from 
51  to  63  horse-power,  in  amounts  which  bear  no  relation  to  the  power  developed. 
This  is  a  common  experience  in  all  friction  tests  of  engines.  It  is  impossible  to  trace 
any  relation  between  the  load  and  the  friction  loss.  More  tests  can  probably  be  quoted 
to  show  that  the  friction  loss  is  constant,  or  nearly  so,  than  to  support  any  other 
supposition. 

"Engineering  Record,  May  28,   1904. 

Since  the  above  was  in  type  a  test  has  been  reported  upon  one  of  the  7,500  horse- 
power, twin,  vertical-horizontal,  Reynolds-Corliss  engines  of  the  Interborough  Rapid 
Transit  Company,  New  York  City.  The  apparent  combined  efficiency  of  engine  and 
generator  was  92.4  per  cent. 

!  Traction  and  Transmission,  London,  January,   1902. 
Technology  Quarterly,  September,   1898. 
Test  by  Prof.  D.  S.  Jacobus. 


180  STEAM  TURBINES 

5.  In  Prof.  Thurston's  papers  upon  the  friction  loss  in  steam  engines, 
previously    referred    to    in    this    chapter,    are    reports    of    several    tests: 
(a)  A  50  horse-power,  Straight  Line,  high-speed  engine,  non-condensing, 
had  a  constant  friction  load  of  about  3  horse-power,  or  6  per  cent  of  full    . 
load,  giving  an  efficiency  of  94  per  cent,     (b)    A  compound,  condensing 
engine  had  a  friction  horse-power  of  44  when  developing  347  horse-power 
and    a    friction    horse-power    of    40    when    developing    185    horse-power. 
(c)    A    16  x  30   Porter-Allen   engine  had   a  friction   horse-power   of   12.7 
when  developing^  142  horse-power  and  8.4  at  84  horse-power.     The  last 
two  engines  (b  and  c)  probably  had  full  load  efficiencies  of  about  87  and 
90  per  cent,  respectively. 

6.  In  addition  to  the  above,   reference  may  be  made  to  the  internal 
friction   of  large   pumping   engines,   which   is   usually  about   10  per   cent. 
The  famous  Leavitt  pumping  engine  at  the  Boston  sewage  works  showed 
an  efficiency  of  90  per  cent,  and  the  Chestnut  Hill,  Mass.,  engine  by  the 
same  designer  had  an  efficiency  of  93  per  cent. 

Summary  of  Engine  Friction  Tests. — Tests  1,  3  and  4  give  the 
combined  efficiency  of  engine  and  generator,  for  three  large  units, 
ranging  from  850  to  5,400  horse-power,  one  of  which  is  vertical. 
These  are  all  modern,  Corliss-type  engines.  Taking  the  two 
horizontal  engines,  the  efficiency  of  No.  3  is  90  per  cent  and  of 
No.  4  93  per  cent,  at  about  normal  load.  The  average  efficiency 
of  No.  4  at  the  different  loads  is  90  per  cent.  It  would  seem  that 
for  large  engines  of  this  type  an  estimate  of  90  per  cent  would  be 
conservative  for  combined  efficiency  at  normal  load.  For  vertical 
engines  the  combined  efficiency  would  be  higher,  reaching  94  per 
cent  (as  a  safe  figure)  in  the  large  sizes. 

In  the  light  of  the  above  efficiency  tests  the  following  table  of 
the  mechanical  efficiency  of  engines  has  been  prepared : 

TABLE  V. 

MECHANICAL  EFFICIENCY  OF  ENGINES  AT  OR  NEAR  THEIR 
NORMAL  LOAD. 


Engine. 

Efficiency,  Per  Cent. 

Engine 
Alone. 

Engine  and 
Generator. 

Large  Vert  ical  Corliss,  Compound  
Large  Horizontal  Corliss,  Compound 

9">.  5  to  97.5 
P4.5 
91 
94 
87     to  93 
90 

92  to  94 
90 

Large  Horizontal,  Triple  

High  Speed  Simple                                  .               ...        ... 

Large  Pumping  

STEAM  TURBINE  PERFORMANCE 


181 


The  Thermal  Unit  Basis  of  Performance. — In  comparing  the 
results  of  engine  and  turbine  tests,  and  especially  where  boiler 
pressures  differ  or  superheated  steam  is  used,  the  efficiencies 
should  be  calculated  on  the  basis  of  the  heat  units  contained  in  the 
steam.  Under  conditions  of  varying  pressure  or  of  superheat  the 
pounds  of  steam  per  horse-power  per  hour  do  not  indicate  the 
amount  of  heat  energy  contained  in  the  steam. 

Calculations  are  given  herewith  to  illustrate  the  heat-unit 
method,  taken  from  a  report  of  tests  upon  a  400  Kw.  Westing- 
house-Parsons  steam  turbine,  by  Messrs.  Dean  and  Main.  The 
efficiencies  for  superheated  steam  are  figured  by  using  0.48  as  the 
value  for  the  specific  heat  of  superheated  steam.  These  calcula- 
tions will  prove  of  assistance  should  the  reader  desire  to  recalculate 
any  tests  on  the  heat-unit  basis. 

CALCULATION  OF  EFFICIENCIES  OF  400  Kw.  TURBINE. 


Dry 

Steam. 

100  Degrees 
Superheat. 

1FO  Degrees 
Superheat. 

593  1? 

594  60 

592  27 

Corresponding     indicated     or     internal 
horse-power  of  a  reciprocating  engine 

-  B.  H.  P.  -j-o.94  ;....;.... 

631  03 

632  55 

630  07 

Total  steam  used  per  hour,  pounds  

8,249 

7  384 

6  779 

Steam  used  per  internal  horse-power  per 

13  07 

11  67 

10  7fi 

Absolute  steam  pressure,  pounds  

168  57 

169  98 

167  84 

Superheat  (exact  figures)  

o 

109  Deg  P 

181  Deg  P 

Temperature  condensed  steam 

95  3  Deg  F 

f)5  8  Deg  F 

Heat  in  one  pound  of  dry  saturated  steam 
at  above  pressures,  B.  T.  U  

1194  1 

1194  3 

1193  9 

Heat  in  superheat  per  pound,  B.  T.  U.  (on 
the  basis  of  specific  heat  —  0.48)  

o 

52  3 

86  9 

Total  heat  in  one  pound  of  steam,  B.  T.  U. 
Heat  of  liquid  in  condensed  steam,  B.  T.  U. 

1194.1 
63.3 

1246.6 
63.8 

1280.8 

62.7 

Heat  used  by  turbine  per  pound,  B.  T.  U. 

1130.8 

1182.8 

1218.1 

From  the  above  the  following  results  are  obtained : 
Case  of  Dry  Steam. 

B.  T.  U.  used  by  turbine  per  minute=(  1,130.8X8,249)  ^60= 
155,466  B.  T.  U. 


182  STEAM  TURBINES 

B.  T.  U.  used  per  internal  horse-power  per  minute,  155,466-=- 
631.03=246.37  B.  T.  U. 

33  000 
Thermal  efficiency^— -     '  =17.22  per  cent. 

U.O  i  /\  I  io 


Case  of  100°  Superheat. 

B.  T.  U.  per  minute,  (1,182X7,384) -^60=145,563  B.  T.  U. 
B.  T.  U.  used  per  internal  horse-power  per  minute, 
145,563-^-632.55=230.12  B.  T.  U. 

33,000 

Thermal  efficiency,—  —=0.1843=18.43  per  cent. 

230.12X778 

Case  of  180°  Superheat. 

B.  T.  U.  per  minute,  (1,218.1X6,779) -=-60=137,625  B.  T.  U. 
B.  T.  U.  used  per  internal  horse-power  per  minute, 
137,563^-630.07=218.33  B.  T.  U. 

33,000 

Thermal  efficiency, =0.1943=19.43  per  cent. 

218.33X778 

Results  of  Turbine  Tests. 

Tables  VI.  to  XVII. ,  inclusive,  contain  results  of  tests  upon 
turbines  of  different  types  and  show  in  condensed  form  what 
economy  may  be  expected  of  turbines  operating  under  different 
conditions. 


STEAM  TURBINE  PERFORMANCE 

TABLE  VI. 
TESTS  ON  30  H.  P.  DE  LAVAL  TURBINE.* 

Non-Condensing.    Jnitial  Pressure  99.54  Ib.  Absolute. 


183 


Half  Load. 

Full  Load. 

Saturated 
Steam. 

Super- 
heated 
Steam. 

Saturated 
Steam. 

Super- 
heated 
Steam. 

Temperature  of  Steam  Degrees  F 
Brake  horse  power  

327 
21.1 
43.3 
2'2 

860 
24.2 
31.5 
588 

327 
43.5 
39.6 
212 

932 
51.2 
25.7 
649 

Lb.  steam  per  B.  H.  P.  per  hour.  . 
Temperature  Exhaust  Degrees  F 

*Made  at  Polytechnical  College,  Dresden. 

TABLE  VII. 
TESTS  ON  DE  LAVAL  TURBINES  AT  DIFFERENT  LOADS.* 


Turbine 
Machine. 

Initial 
Pressure. 
Lb.  per 
sq.  inch. 

Vacuum, 
inches. 

No.  of 
nozzles 
open. 

Electrical 
H.  P. 

Lb.  steam 
per 
electrical 
H.  P. 
per  hour. 

Remarks. 

103.7 

25.8 

5 

92.7 

22.6 

100  H.  P.  Turbine 

103.8 

26.4 

3 

55.6 

22.7 

Saturated 

Dynamo. 

107.4 
106.7 

26.8 
27.9 

2 

1 

35.0 
15.5 

24.7 

27.8 

Steam. 

Lb.  of 

steam 

Brake 

per 

H.  P. 

Brake 

H.  P. 

per  hour. 

113.8 

26.4 

7 

163.0 

17.6 

116.9 

25.9 

6 

138.4 

18.2 

150  H.  P.  Turbine 

113.8 

26.2 

5 

114.5 

17.9 

Saturated 

Motor. 

114.3 

26.5 

4 

83.3 

18.7 

Steam. 

Hi.  4 

27.0 

3 

64.1 

19.0 

116.,! 

25.7 

2 

37.5 

2,>.3 

192.7 

27.3 

7 

303.6 

14.1 

196.3 

27.6 

6 

255.5 

14.7 

Steam 

300  H.  P.  Turbine 
Motor. 

196.3 
196.3 
190.6 
196.3 

27.6 
27.6 

27.8 
28.1 

5 
4 
3 
2 

216.9 
172.6 
121.6 

74.2 

14.4 
14.5 
14.9 
17.2 

Super- 
heated 60 
degrees  F. 

413.3 

28.5 

1 

31.5 

21.6 

*From  paper  by  Konrad  Anderson  in  Trans.  Inst.  Eng.  and  Shipbuilders  in  Scot- 
land, Nov.,  1902. 


184 


STEAM  TURBINES 


TABLE  VIII. 

TESTS  ON  300  H.  P.  DE  LAVAL  TURBINE  DYNAMO. 
(DEAN  AND  MAIN.) 


Pressures 

Steam  Used 

Lb.  sq.  in.  gauge. 

per  Hour. 

Number. 

Superheat, 
Dtgrees  F. 

Vacuum, 
Inches. 

Brake 
Horse 
Power. 

Above 
Governor 
Valve. 

Below 
Governor 
Valve. 

Total 
Pounds. 

Pounds 
B.  B^P. 

Tests  with 

Saturated 

Steam  -Av 

erage    Res 

ults. 

1 

206.4 

196.9 

26.6 

ass 

5052 

15.17 

2 

207.3 

196.5 

26.8 

284.8 

4430 

15.56 

3 

207.6 

195.8 

W.86 

195.2 

3229 

16.54 

4 

201.5 

197.9 

.... 

28.1 

118.9 

1950 

16.40 

T 

ests  with  S 

uperheate 

d    Steam  - 

Average  R 

esults. 

5 

207.0 

198.5 

84.0 

27.2 

a52.0 

4906 

13.94 

6 

207.4 

197.0 

64.0 

27.4 

298.4 

4:282 

14.35 

7 

200.7 

196.6 

10.0 

27.5 

196.5 

3033 

15.44 

8 

202.4 

197.7 

16.0 

2?.  4 

196.5 

3062 

15.62 

9 

Average 

of  7  and  8. 

15.53 

TABLE  IX. 

SUMMARY  OF  DEAN  AND  MAIN  TESTS  ON  300  H.  P.  DE  LAVAL  TURBINE. 
Relative  Steam  Consumption  at  Different  Loads. 


Group  No. 

Loads 
B.  H.  P. 

Relative 
Loads. 

Steam  per 
Brake 
Horse  Power. 

Increase  for 
Diminishing 
Loads,  referred  to 
Maximum  Load. 

Saturated  Stea 

tn. 

2 
3 
4 

333 

2S5 
195 
119 

10W 
86$ 
59* 
36* 

15.171b. 
15.56  " 
16.54  " 
16.40  " 

2.6* 
9.0* 

8.1* 

Superheated  St 

earn 

5 
6 
9 

352 
298 
196 

100* 

85?, 
6# 

13.941b. 
14.35  " 
15.53  u 

2.9* 
11.4* 

Saving  (at  the  Turbine)  by  the  Use  of  Superheated  Steam. 


Amount  of 
Supei-heat. 

Load  with 
Superheat- 
ed Steam. 

Load  with 
Saturated 
Steam. 

Steam 
used  per 
Brake  H.  P. 

with  Sup. 
Steam. 

Dry  Steam 
used  per 
Brake  H.  P. 
with  Sat. 
Steam. 

Saving  by 
use  of 
Superheat- 
ed Steam. 

1  and  5 
2  and  6 

84°  F. 
64°  F. 

352  H.  P. 

298      " 

333  H.  P. 

285      " 

13.94  Ib. 
14.35  " 

15.171b 
15.56  " 

8.8* 
8.4* 

STEAM  TURBINE  PERFORMANCE 


185 


TABLE  X. 
MISCELLANEOUS  TESTS  ON  PARSONS  TURBINES.* 


At  Stop  Valve 

Vacuum, 
Inches. 

Speed, 
Rev.  per 
Minute. 

Load  in 
Kw. 

Steam  Used  per  Hour. 

Pressure 
Above 
Atmosphere, 
Ib.  per  sq  in. 

Superheat, 
Degrees  F. 

Total 
Pounds. 

Pounds 
per  Kw. 

Test 

No.  /.     7J-A' 

w.  Continu 

ous  -  Curre 

nt  Turbo  f 

or  Banburv. 

141.2 
144 
142 

64.2 
0 
0 

27.1 
27.0 
27.1 

4,140 
4,140 
4,140 

75.7 
75.2 
56.6 

2,006 
2,201 
1,777 

26.4 
29.2 
31.2 

Test 

No.  2.    zoo-K 

w.  Continu 

ous  -  Curre 

nt  Turbo  f 

or  Shipley 

150 

151 
156 
151 

57 
55 
181 
166 

27 
27.9 
2;.  3 
28.0 

3,000 
3,000 
3,000 
3,000 

204.2 
101.2 
202.5 

100.27 

4,538 
2,698 
4,130 
2,446 

22.23 
26.67 
20.39 
24.41 

Te 

stNo.3.    37 

5-Kw.  Tur 

bo-Alterna 

tor  for  Du 

ndee. 

152.9 
149.4 

148!9 

27.4 
27.5 

3,000 
3,000 

376.9 
374.06 

8,143 
7,202 

21.6 
19.25 

Test 

No.  4.  300-  K 

w.    Turbo  - 

Alternate 

r.—Hulton 

Colliery. 

158.0 
157.0 
152.0 
154.0 
158.0 

0 
0 
0 
0 
0 

15.33 
19.33 
22.33 
25.33 
26.58 

3,000 
3,000 
3,000 
3,000 
3,000 

297.4 
305.1 
303.4 
303.15 
303.2 

8,732 
8,369 
7,764 
7,336 
7,020 

29.36 
27.43 
25.59 
24.19 
23.15 

Test  No.  3. 

300-  Kw.  Tu 

rbo-Altern 

ator.  —  De 

Beers  Exp 

losive  Works 

150.0 
153.0 
150.5 

53.3 
50.0 
40.2 

27.88 
27.78 
27.9 

3,000 
3,000 
3,000 

312.1 
231.8 
1'4.5 

6,260 
4,960 
3,670 

20.06 
21.45 
23.75 

Test  No.  6. 

i^oo-Kw.  T 

urbo-Alter 

nator.  —  Ne 

w  castle  -  on 

-TyneE.S.  C 

0. 

196 
197 
196 
199 
200 

Af 

76 
84 
76 

77 
68 

ter  16  month 

27.45 
27.35 
27.95 
28.35 
28.45 

s'  use  -the 

1,200 
1,200 
1,200 
1,200 
1,200 

following 

1,442 
1,015.5 
714.0 
360.5 

25,962 
20,124 
15,288 
9,114 
2,948 

re  obtained: 

18.0 
19.8 
21.4 
25.2 

figures  we 

203 
207 

92 
60 

26.11 
26.46 

1,210 
1,208 

1,823 
1,513 

32,431 

27,582 

17.7 
18.23 

Test 

No.  7.    /v5"oo- 

Kw.  Turbo 

-Alternate 

rfor  Sheffi 

eld  Corporat 

ion. 

With  vacu 

um  augment 

or,  and  inc 

luding  450 

Ibs.  steam 

per  hour  us 

ed  by  it. 

113.6 
111.6 
141 
154 

108.3 
156.4 
113 
47.5 

26.69 
27.12 
27.72 

27.72 

1,455 
1,500 
1,500 
1,500 

1,316.5 
1,061.6 
512.7 
0 

2t,732 
19,830 
11,425 
3,1:28 

18.76 
18.66 
22.3 
0 

*From  a  paper  by  Hon.  Chas.  A.  Parsons,  G.  Gerald  Storey,  and  C.  P  Martin,  pre- 
sented before  the  British  Institute  of  Electrical  En  jineers,  May,  1904.  In  summariz- 
ing the  tests  of  this  table,  and  other  tests  upon  his  turbines,  Mr.  Parsons  states:  "  It 
will  be  seen  that  under  conditions  of,  say,  HO  pounds  steam  pressure  and  100  degrees 
superheat,  and  a  vacuum  of  27  inches,  the  consumptions  in  round  numbers  are  as  fol- 
lows: A  100  Kw.  plant  takes  about  25  pounds  of  steam  per  Kw.  hour  at  full  load;  a  200 
Kw.  takes  22  pounds;  a  500  Kw.  20  pounds;  a  1,000  Kw.  19  pounds;  a  1,500  Kw.  18  pounds; 
and  a  3,000  Kw.  16  pounds.  These  figures  are  derived  from  averages  of  a  large  number 
of  tests  that  have  been  made  from  time  to  time.  Without  superheat  the  consumptions 
are  about  10  per  cent  more." 


186 


STEAM  TURBINES 
TABLE  X.    (CONTINUED), 


At  Stop  Valve. 

Steam  Used  per  Hour. 

Pressure 
Above 

Superheat, 

Vacuum, 
inches.  - 

Speed, 
Rev.  per 
Minute. 

Load  in 
Kw. 

Total 

Pounds 

Atmosphere, 

Degrees  F. 

Pounds. 

per  Kw. 

Ib.  per  sq.  in. 

W 

ithout  vac 

uum  augm 

entor. 

115.6 

143 

25.18 

1,500 

1,029  3 

21,264 

20.7 

137 

119 

25.97 

1,500 

534.25 

12,820 

24.02 

150.3 

72.4 

2o.62 

1,500 

0 

2,957.4 

0 

3,ooo-Kw. 

Parsons  Tu 

rbo-Altern 

ator  Suppl 

fed  to  Fra 

nkjort  by  Me 

ssrs. 

Brown, 

B  overt  & 

Co. 

138.5 

235 

27 

1,350 

2,993 

44.200 

14.74 

170.5 

187 

27.5 

1,350 

2,518 

38,300 

15.59 

142 

120 

27.2 

1,350 

2,600 

41,200 

15.8 

139 

114 

27.2 

1,350 

2,600 

41,400 

15.9 

168.5 

184 

27.9 

1,350 

1,945 

30,800 

15.84 

146 

120 

27.6 

1,350 

2,000 

32,600 

16.3 

137 

101 

27.4 

1,350 

1,442 

25,400 

17.6 

142 

30 

29.3 

1,350 

0 

4,700 

excited 

142 

30 

29.3 

1,350 

0 

3,560 

non-exci'ed 

TABLE  No.   XT. 
TESTS  ON  PARSONS  TURBINES,  WHEN  RUNNING  NON-CONDENSING.* 


At  Stop  Valve. 

Back 
Pressure, 
Lb.  per  Sq. 
In.  Gauge 

Vacuum, 
Inches. 

Speed, 
Rev.  per 
Minute. 

Load  in 
Kw. 

Steam  Used  per  Hour. 

Pressure 
Above 
Atmos- 
phere, Lb 
per  Sq.  In 

Super- 
heat, 
Degrees 

F. 

Total 
Pounds 

Pounds 
per  Kw. 

Test  N 

o.  i.    250-  K\ 

w.  Continu 

ous-Curr 

ent.  —Me 

ssrs.  Gut 

nness,    Son 

&°Co. 

144 
142.6 
138 
143 

0 
0 
0 
0 

0 
6 
11.1 
11 

'.'.'.'.'. 

3,047 
3,047 
3,055 
3,115 

251.55 
255.82 
253.15 
125.45 

9,510 
10,584 
11,  K4 
7,4;  5 

37.80 
41  38 
44.15 
59.58 

Test  No. 

2.  joo-Kw. 

Turbo-A 

Iternato 

r.  —Hult 

on  Colliery. 

161 
158 

0 
0 

:::: 

0 

26.58 

3,040 
3,000 

296.6 
303.2 

10,180 
7,020 

34.2 
23.15 

Te 

st  No.  3.    3 

oo-Kw.   Tu 

rbo  -  Alte 

rnator.— 

Metropol 

itan  E.  C.  C 

0. 

142 
146 
146 

0 
0  - 
0 

.... 

0 
22.57 

28. 

1,800 
1,800 
1,800 

506.2 
509.85 
500 

16,903 
13,714 

33.39 

26.89 
23.  5t 

*From  a  paper  by  Hon.  Charles  A.  Parsons,  G.  Gerald  Storey  and  C.  P.  Martin, 
presented  before  the  British  Institute  of  Electrical  Engineers,  May,  19l4. 
tResult  estimated  by  the  author  from  curve  plotted  in  original  paper. 


STEAM  TURBINE  PERFORMANCE 

TABLE  XII. 

TESTS  ON  RATEAU  TURBINE  OF  500  H.  P.  LOCATED  AT 
PENARROYA,  SPAIN.* 


187 


Overload 

Data. 

V*  Load. 

%  Load. 

Full 
Load. 

Over- 
load. 

at  in- 
creased 
speed 

- 

(2400  rev.) 

Electrical  H.  P.  at  brushes.  .  . 

135 

259 

525 

627 

641 

Admission  pressure,absolute, 

Ib  per  sq.  in  

46.21 

76.6 

136 

156 

156 

Exhaust  pressure,    absolute. 

Ib  per  sq.  in  

1.24 

1.33 

1.63 

1.82 

1.82 

Theoretical  steam  consump- 

tion of  perfect    engine  per 

H  P  hour  Ib 

10.93 

9.8 

8.89 

8.73 

8.73 

Actual     steam    consumption 

Eer  electrical  H.  P.  hour  at 

rushes,  Ib  

21.3 

18 

15.8 

15.39 

14.90 

Combined    efficiency    of   the 

electrical  generating  set  re- 
ferred to  the  perfect  engine 

0.513 

0.540 

0.560 

0.569 

0.580 

TABLE  XIII. 
TESTS  ON  A  ZOELLY  TURBINE  OF  500  BRAKE  HORSE  POWER. f 


0) 

In  Steam  Pipe. 

After  Passing 
Governor  Valve. 

Steam  Used  per 
Hour. 

| 

Net 

£ 

Vacuum, 

Power 
• 

fe 

« 

D 

Pressure 
Ib.  per 
sq.  in. 

Super- 
heat 
Degrees 

Pressure 
Ib.  per 
sq.  in. 

Super- 
heat 
Degrees 

Inches. 

in 
Kilo- 
watts. 

Total 
Pounds. 

Pounds 

£ 

Absolute. 

P. 

Absolute. 

P. 

T 

ests  with 

Dry,  Sat 

urated  S 

team. 

i 

158.47 

.... 

143.3 

363.06 

7903 

21.75 

2 

158.47 

143.4 

28.5 

387.65 

8327 

21.50 

3 

154.98 

128.4 

28.5 

334.51 

7825 

22.21 

4 

156.34 

\ 

98.2 

\ 

28.4 

240.1 

5778 

24.08 

5 

155.77 

77.4 

28.3 

182.22 

4683 

25.70 

6 

156.76 

\ 

43.5 

28.6 

80.13 

2650 

33.08 

7 

156.62 

17.3 

39  4 

28.6 

1025 

8 

158.89 

. 

10.6 

53!  o 

28.5 



649 



Tests  with 

Superhe 

ated   Ste 

am. 

9 

181.9 

135 

138.0 

102 

28.7 

391.66 

7454 

19.08 

10 

186.44 

153 

107 

28.7 

389.6 

7335 

18.82 

11 

159.89 

109 

139  !o 

101 

28.7 

390.4 

7730 

19.90 

*Quoted  by  Prof.  Rateau  in  various  papers,  including  one  presented  before  the 
A.  S.  M.  E.  in  June,  1904. 

tFrom  Data  furnished  the  Author  by  the  Builders. 


188 


STEAM  TURBINES 


TABLE  XIV. 

TESTS  ON  500  Kw.  CURTIS  TURBINE,  CORK  (IRELAND)  ELECTRIC 
TRAMWAY  AND  LIGHTING  Co.'s  STATION.* 


Initial 
Pressure 
Ib.  sq.  in. 
Gauge. 

Super- 
heat 
Degrees 
F. 

Vacuum, 
Inches. 

Average 
Load 
in  Kw. 

Approx. 
Load. 

Steam  Used  per  Hour. 

Total 
Pounds. 

Pounds 
per  Kw. 

155 

51 

28.8 

125.87 

3253 

25.9 

155 

50 

28.6 

250.06 

1A 

5600 

22.4 

153 

70 

27.8 

393.8 

JJ 

8251 

20.9 

153 

104 

26.9 

511.7 

Full 

10503 

20.5 

151 

124 

26.2 

613.5 

1V4 

1-2811 

20.9 

* Electrical  Review  (English),  Nov.  18,  1904. 


TABLE  XV. 

TESTS  ON  2oco  Kw.  CURTIS  TURBINE.* 


Load  in 
Kw. 

Revolu- 
tions per 
Minute. 

Initial 
Pressure, 
Gauge. 

Vacuum, 
Referred  to 
30  in.  Merc. 

Superheat, 
Deg.  F. 

Lb.  Steam  per 
Kw.  Hour. 

1970 

918 

162 

28.15 

210 

15.12 

1040 

928 

167 

28.38 

190 

15.87 

560 

930 

163.5 

28.40 

210 

17.86 

2005 

918 

169 

28.37 

125 

15.80 

1066 

928 

171 

28.48 

105 

16.33 

1970 
0 

918 
932 

165 
165.5 

28.21 
28.05 

105 
157 

16.20 

Steam  per  Hour. 
1530 

*Test  by  A.  R.  Dodge,  Schenectady,  N.  Y.     Reported  by  August  H.  Kruesi,  in  a 
paper  before  the  National  Electric  Light  Association,  June,  1905. 


STEAM  TURBINE  PERFORMANCE 


189 


TABLE  XVI. 
TESTS  ON  400  Kw.  (580  H.  P.)  WESTINGHOUSE-PARSONS  TURBINE.* 


Initial 
Pressure, 
Ib.  sq.  in. 
Gauge. 

Super- 
heat, 
Degrees 
F. 

Vacuum, 
Inches. 

Brake 
Horse 
Power. 

Steam  Used 
per  Hour. 

Total 
Pounds. 

Pounds 
B.?!rp. 

Resu 
3\%  Overload  

Its  with  A 

150 
156 
154 
153 

.ts  with  A 

151 
154 

esults  with 

153 
154 
156 
156 

ts  with  D 

152' 
155 

pproxim 

104 
109 
104 

87 

pproxim 

182 
181 

Dry,  Sat 
ry  Steam 

ately    100° 

27.10 
27.06 
27.10 
27.10 

ately    180° 

27.00 
27.10 

urated  Ste 

26.87 
26.84 
26.80 
26.90 

and  Poor 

25.90 
25.91 

F.  Super 

758.9 
594.6 
445.3 
239.9 

F.  Super 

762.6 
592.3 

am. 

728.4 
593.2 
448.0 
241.3 

Vacuum. 

694.8 
593.1 

heat. 

9157 
7384 
5728 
3500 

heat. 

8520 
6779 

9928 
8249 
6486 
3876 

9734 
8514 

12.07 
12.41 
12.86 
14.62 

11.17 
11.45 

13.63 
13.91 
14.48 
16.06 

14.01 
14.35 

Full  Load  (2%  Overload).  . 
77%  Load  

41%  Load 

Resu 
32%  Overload 

Full  Load  (2%  Overload).  . 
R 
20%  Overload  

Full  Load  (2%  Overload).  . 
77%  Load  

42%  Load  

Resu 
20%  Overload  

Full  Load  (2%  Overload).  . 

With  100°  Superheat  the  speed  of  rotation  was  3,547  R.  P.  M.  at  full  load;  at  31j< 
overload  it  dropped  2.5%  and  at  41%  load  it  increased  1.2%. 

With  Dry  Steam  the  speed  was  3,545  R.  P.  M.  at  full  load;  at  26%  overload  it  drop- 
ped 1.8%  and  at  42%  load  it  increased  1.6%.  With  turbine  running  light  speed  increased 

3.4*. 


*Tests  made  at  the  works  of  the  builders  by  Dean  and  Main  in  1903. 


190 


STEAM  TURBINES 
TABLE  XVII. 


TESTS  ON  A  1250  Kw.  TURBINE  FOR  INTERBOROUGH  RAPID 
TRANSIT  Co.,  NEW  YORK.* 


w 

Load  Carried  by  Turbine. 

Steam  Used  per  Hour. 

2J  bo 

8 

3  3 

•S 

so 

_j  . 

a 

1 

*o 

«:' 

~  • 

0>  03 

•gi 

B 

Id 

iLn- 

c  * 

Is 

«j 

T3 

W 

In 

<U 

J 

|sr 

II 

s 
§ 

W  £ 

M  "*  > 

K 

2§ 

ft 

P'ofe' 

0,0 

coQ 

5 

£* 

£ffi£ 

£fo 

££ 

jS 

SSa 

Re 

suits  w 

ith  Dry, 

Saturate 

d  Stea 

m. 

'     150.3 

27.08 

196.95 

264.0 

0.157 

7155.0 

36.32 

H27.08 

151.4 

27.11 

342.73 

459.42 

0.276 

9732.0 

2H.4 

21.18 

146.8 

27.1 

655.98 

879.33 

0.525 

15074.0 

22.98 

17.14 

27"              146.5 

\ 

27.11 

939.53 

1326.5 

0.790 

20254.0 

20.46 

15.27 

Vacuum         147.1 

27.11 

1321.46 

1771.4 

1.060 

25712.0 

19.47 

14.52 

148.0 

27.05 

1489.4 

1996.5 

1.190 

28208.0 

18.95 

14.13 

144.5 

27.05 

1713.5 

2297.1 

1.3"0 

330oG.O 

19.28 

14.38 

141.8 

26.79 

1983.9 

2666.0 

1.590 

40547.0 

20.39 

15.21 

23/,               151.0 

28.05 

334.78 

448.74 

0.268 

9295.0 

27.78 

20.71 

TT                       146.8 
Vacuum          146.1 

28.1 
28.08 

972.0 
1363.95 

1303.0 
1828.3 

0.778 
1.091 

19334.0 
25639.0 

19.99 
18.79 

14.91 
14.02 

R 

esuits 

with  75° 

F.  Super 

heat. 

f     151.8 

76.6 

27.15 

191.0 

256.0 

0.153 

6734.0 

35.25 

26.3 

a-"              151.9 

77.2 

27.07 

333.55 

447.1 

0.267 

9170.0 

27.58 

20.51 

,T      '           J      150.0 
Vacuum   ]      147-7 

77.0 
76.15 

27.15 
27.1 

664.67 
986.23 

891.0 
1322.04 

0.531 
0.790 

14181.0 
19108.0 

21.33 
19.39 

15.9 
14.45 

1      116.  3 

76.0 

27.1 

1293.9 

1734.9 

1.038 

23903.0 

18.48 

13.78 

f     151.8 

75.5 

23.05 

198.4 

266.1 

0.159 

6300.0 

31.76 

23.68 

28"              150.8 

77.4 

28.05 

333.15 

446.6 

0.266 

8420.0 

25.39 

18.86 

Vacuum    i      147.6 

76.5 

28.1 

977.64 

1310.5 

0.780 

18180.0 

18.59 

13.87 

[     146.0 

78.25 

28.1 

1274.2 

1708.0 

1.020 

22504.0 

17.66 

13.17 

Tests  upon  a  500  Kw.  Curtis  Turbine  at  Newport,  R.  I. — A 
Curtis  turbine  at  the  Newport,  R.  I.,  power  house  of  the  Old 
Colony  Street  Railway  Company,  a  description  of  which  plant  will 
be  given  later,  was  tested  by  George  H.  Barrus,  consulting  en- 
gineer, Boston,  Mass.f  This  turbine  is  one  of  the  earlier  two- 
stage  type  and  had  been  in  continuous  service  nearly  a  year  when 
the  test  was  made.  The  results  with  dry,  saturated  steam,  150 
pounds  initial  pressure,  and  two  inches  back  pressure  in  the  con- 
denser, are  tabulated  herewith  for  different  loads,  together 

*Tests  made  at  the  Westinghouse  Machine  Company's  Plant,  and  reported  by  A.  M. 
Mattice,  Chief  Engineer. 

+The  Iron  Age,  May  5,  1904. 


STEAM  TURBINE  PERFORMANCE  191 

with  some  general  results  with  both  saturated  and  super- 
heated steam. 

At  full  load,  19.78  pounds 

At  three-fourths  load,  20.69       " 

At  half  load,  21.38 

At  one-fourth  load,  27.85 

At  50  per  cent  overload,  20.22       " 

With  the  variable  commercial  load  at  the  station,  which  ranges 
from  333  Kw.  to  114  Kw.  and  averaged  253.2  Kw.,  the  con- 
sumption of  dry  steam  was  22.38  pounds  per  Kw.  hour.  When 
the  commercial  load  was  augmented  by  a  constant  rheostat  load, 
bringing  the  average  up  to  421.9  Kw.,  the  steam  consumption  was 
20.7  pounds  per  Kw.  hour. 

With  superheating  of  150.5  degrees  at  the  throttle  valve,  the 
steam  consumption  was  17.79  pounds  per  Kw.  hour  at  full  load, 
or  10  per  cent  less  than  with  saturated  steam.  When  superheat- 
ing was  increased  to  289.6  degrees,  the  consumption  was  15.1 
pounds  per  Kw.  hour,  or  19.6  per  cent  less  than  with  saturated 
steam.  In  these  tests  the  load  on  the  auxiliaries  averaged  about 
16.7  Kw.,  or  about  3  per  cent  of  the  normal  rating  of  the  turbine. 

Best  Turbine  Results. — Table  XVIII.  summarizes  the  best  re- 
sults of  tests  upon  turbines  quoted  in  this  chapter.  With  the  ex- 
ception of  tests  upon  two  turbines,  these  results  are  all  in  terms  of 
kilowatts  or  electrical  horse-power,  and  before  any  comparisons 
can  be  made  with  piston  engines,  there  must  be  allowances  for 
efficiency,  as  explained  in  the  first  part  of  the  chapter.  To 
facilitate  comparisons,  certain  efficiencies  have  been  assumed,  as 
given  in  the  last  column  of  the  table;  and  by  using  these  effi- 
ciencies the  equivalent  rates  of  consumption  per  indicated  horse- 
power per  hour  were  calculated  and  tabulated  in  Column  6.  The 
efficiencies  were  chosen  on  the  following  basis : 

Combined  efficiency  of  generator  and  engine  in  units  of  from 
400  to  500  H.  P.  (or  300  to  400  Kw.),  88  per  cent. 

Combined  efficiency  of  units  of  from  500  to  1,500  Kw.,  90  per 
cent. 

Combined  efficiency  of  2,000  to  3,000  Kw.  units,  92  per  cent. 


192  STEAM  TURBINES 

TABLE  XVIII. 

EXAMPLES  OF  STEAM  CONSUMPTION  OF  TURBINES.     BEST  RESULTS 
OF  TESTS  QUOTED  IN  THIS  CHAPTER. 


Turbine. 

Nominal 
Power. 

Steam  Used  per  Hour. 

Estimated 
Equival'nt 
Consump- 
tion per 
I.  H.  P. 

Per  cent 
efficiency 
assumed 
in  estimat- 
ing I.  H.  P. 
Results. 

Pounds 
per 
B.  H.  P. 

Pounds 
per 
E.  H.  P. 

Pounds 
per  Kw. 

| 

2 

3 

4 

5 

6 

7 

Results 

with  Sat 

urated  St 

earn. 

De  Laval 

300  H  P 

15  17 

13  96 

92 

Rateau  

500  H.  P. 

14  90 



13  11 

88 

Zoellv 

500  H.  P. 

16  05 

21  50 

14  12 

88 

Curtis  (American)  

500  Kw. 
400  Kw. 
1250  Kw. 

is!  63 

14.76 
i4!l3 

19.78 

is!95 

13.28 
12.68 
12.72 

90 
93 

90 

Westinghouse  -  Parsons 
Westinghouse  -  Parsons 

Results  with  Su 

perheated 

Steam(M 

oderate 

Superhe 

at,  not 

e 

xceeding 

150°). 

De  Laval 

300  H   P 

13  94 

12  82 

92 

Zoelly  

500  H.  P. 

14  05 

18  82 

12  36 

88 

Curtis  (English)  

500  Kw. 
500  Kw. 

15.29 
13.28 

20.50 
17  79 

13.76 
11  95 

90 
90 

Curtis  (American)  

Parsons           .        . 

300  Kw. 

14  96 

20  06 

13  16 

88 

Parsons  

1500  Kw. 

13.44 

18.0 

12.10 

90 

Parsons         

3000  Kw. 

11  79 

15  8 

10  85 

92 

Westinghouse  -  Parsons 
Westinghouse-  Parsons 

400  Kw. 
1250  Kw. 

12  07 

isi-rs 

is!48 

11.23 
12.40 

93 

90 

Results  with  Hig 

hly  Superh 

eated  Ste 

am(Supe 

rheat  fro 

m  180°  to  2 

90°). 

Curtis  (American)  

500  Kw. 

11.26 

15.1 

10.14 

90 

Curtis  (American)  

2000  Kw. 

11.27 

15.12 

10.36 

92 

Parsons  

3000  Kw. 

11.00 

14.74 

10.12 

92 

Westinghouse-  Parsons 

400  Kw. 

11  17 

10.39 

93 

Engine  efficiency  alone  (without  generator)  for  units  of  from 
400  to  500  H.  P  (or  300  to  400  Kw.),  93  per  cent. 

Engine  efficiency  corresoonding  to  the  300  H.  P.  De  Laval 
turbine,  92  per  cent. 

Chart  for  Estimating  Rate  of  Steam  Consumption. — If  results 
on  any  other  efficiency  basis  are  desired,  they  may  be  easily  cal- 
culated, or  they  may  be  obtained  by  the  aid  of  the  accompanying 
chart,  Fig.  1.  On  this  chart  the  figures  at  the  left  are  pounds  of 
steam  per  electrical  horse-power  per  hour.  The  inclined  lines  are 
iorr various  efficiencies,  and  at  the  top  are  corresponding  values 
for  pounds  of  steam  per  indicated  or  brake  horse-power  per  hour. 
Obviously,  also,  if  brake  horse-power  units  are  assumed  to  be  at 


STEAM  TURBINE  PERFORMANCE 


193 


to 


Pounds  Steam  per  Indicated  or  Brake  Horse  Power  per  Hour. 
8  9  10          11  12  13          14  15          16          17  18          19 


20 


25          21          23          22          21          20  19          18          17 

Pounds  Steam  per  Kilowatt  Hour. 


16 


15          14 


Fig.   1.     Chart  for  Estimating  Rate  of  Steam  Consumption. 

the  left,  indicated  horse-power  units  corresponding  can  be  found 
at  the  top.  The  inclined  line  labeled  "kilowatts  to  horse-power" 
is  for  the  purpose  of  converting  kilowatt  units  at  the  bottom  to 
horse-power  units  at  the  left. 

Example  to  Illustrate  Use  of  Chart.  —  Following  the  dotted 
line,  we  find  that  20  pounds  of  steam  per  kilowatt  hour=14.9 
pounds  per  electrical  horse-power  per  hour.  Assuming  90  per 
cent  efficiency,  the  equivalent  steam  consumption  per  indicated 
horse-power  per  hour  is  found  by  retracing  the  dotted  line  from 
the  14.9  point  until  it  meets  the  90  per  cent  line;  then  extending 
upward  until  it  reaches  the  13.4  point,  which  gives  the  required 
rate  of  consumption. 


194 


STEAM  TURBINES 
TABLE  XIX. 


EXAMPLES  OF  TESTS  UPON  RECIPROCATING  ENGINES  OF  EXCEPTIONALLY 
HIGH  ECONOMY,  SHOWING  BEST  RESULTS  OBTAINED. 


te    . 

tjfo 

S| 

Engine. 

gjj 

«T 

sis, 

i> 

|s 

cd 

0>    Cfl 

II 

sfl 

Authority. 

J-i    £ 

*^  (5  ^ 

o  *? 

>  ,_, 

^    p^  *-j 

££ 

££3 

>£ 

c?Q 

*•*       '  ^ 

Westinghouse  Vertical 

at  Brooklyn,  N.  Y.... 
Rockwood-Wheelock 

5100 

185 

27.3 

76 

.... 

11.93 

Eng.  Record,  May  28, 
1W4 

at  Natick,  R.  I  

595 

159 

25.4 

76.4 

.... 

13 

F.   W.   Dean,    Trans. 

Mclntosh  and  Seymour 
at  Webster,  Mass  

1076 

123 

27.10 

99.6 

20 

12.76 

A.  S.  M.  E.,  1895. 
F.   W.    Dean,    Trans. 

Rice    and    Sargent    at 

A.  S.  M.  E.,  1898. 

Brooklyn 

627 

151 

28.6 

121 

12.10 

D.  W.  Tacobus,  1  'rans. 

Rice    and    Sargent    at 

A.S.M.  E.,  1903. 

Philadelphia      

420 

142 

25.8 

102 

297 

9.56 

D.  W.Tacobus,  Trans. 

Horizontal,  Four-valve 
Leavitt    Pumping   En- 

658 

150.4 

26.4 

80 

16.4 

12.03 

Barrus'  Engine  Tests. 

gine  at  Chestnut  Hill. 

Mass  

575.7 

175.7 

27.25 

50.6 

11.2 

E.  F.  Miller  in    Tech- 

nology   Quarterly, 
Vol   IX. 

Best  Reciprocating  Engine  Performance. — In  Table  XIX.  are 
a  few  "best  results"  selected  from  tests  upon  several  very  economi- 
cal engines.  Many  other  results  as  good  as  these  could  have 
been  tabulated,  but  the  ones  given  are  indicative  of  what  is  now 
attained  under  the  most  favorable  conditions.  It  is  conservative 
to  say  that  compound  engines  may  now  be  built  to  produce  an 
indicated  horse-power  on  12.5  pounds  of  steam  per  hour,  with 
saturated  steam.  With  a  high  degree  of  superheat  the  long- 
sought  10-pound  mark  has  nominally  been  passed,  but  if  the  re- 
_sults  were  recalculated  in  terms  of  the  equivalent  rate  of  con- 
sumption of  saturated  steam,  by  using  the  heat  unit  method,  it 
would  be  found  that  they  barely  reached  10  pounds. 

Average  Engine  Performance. — Since  the  results  in  Table  XIX. 
are  from  picked  tests,  and  are  exceptional,  Table  XX.  is  given, 
which  fairly  represents  what  the  ordinary  high-grade  engine  will 
do.  This  table  is  made  up  from  the  results  of  tests  upon  four- 
valve,  compound  condensing  engines,  published  in  Barrus'  En- 
gine Tests.  They  are  not  selected  tests,  but  are  from  23  engines 
in  commercial  operation  and  are  average  representatives  of  their 
class.  We  find  that  one  of  the  results  falls  below  1:2  pounds,  five 


STEAM  TURBINE  PERFORMANCE 
TABLE  XX. 


195 


RESULTS  OF  TESTS  ON  FOUR-VALVE  ENGINES — THE  FIRST  FIVE  ENGINES 

OPERATED  WITH  STEAM  SLIGHTLY  SUPERHEATED  (12  TO  44.5 

LKGREES);  THE  OTHERS  WITH  SATURATED  STEAM 


Indicated 

Steam  per 

Indicated 

Steam  per 

Horse 

I.  H.  P.  per 

Horse 

1.  H.  P.  per 

Power. 

Hour,  Pounds. 

Power. 

Hour,  Pounds. 

659 

11.89 

725 

13.27 

658 

12.03 

1714 

13.27 

670 

1^.29 

1030 

13.21 

798 

13.28 

843 

13.53 

1017 

13.26 

382 

14.05 

689 

12.69 

373 

14.18 

708 

12.45 

ere 

14.6 

636 

13.28 

1540 

14.1 

20 

13.37 

800 

15.78 

719 

13.09 

606 

16.28 

741 

13.23 

716 

19.36 

729 

13.01 

results  fall  below  13  pounds,  and  16,  or  over  two  thirds,  fall  be- 
Icw  14  pounds,  while  only  three  are  above  14  pounds.  This  table 
confirms  what  is  commonly  accepted  among  engineers ;  viz.,  that 
a  rate  of  consumption  of  from  13  to  14  pounds  is  to  be  expected 
with  engines  of  this  class.  A  figure  of  13.5  pounds  may  be  set 
as  the  rate  for  large  engines,  operating .  under  good  conditions, 
with  saturated,  or  slightly  superheated  steam. 

TABLE  XXL 

TESTS  ON  COMPOUND  ENGINE  AT  GHENT,*  BELGIUM,  BY   PROF.  SCHROETER, 
SHOWING  EFFECT  ON  ECONOMY  OF  VARYING  SUPERHEAT. 


Indicated 
Horse 
Power. 

Superheat, 
Degrees  F. 

Steam  per 
I.  H.  P.  per 
Hour,  Pounds 

Equivalent 
Pounds  of 
Saturated 
Steam. 

H^-at  Units 
per  I.  H.  P. 
per  Hour. 

222 
226 
227 
223 
223 
218 

0 
43.7 
97.7 
151.7 
221.2 
310.9 

13.08 
11.58 
11.00 
10.67 
9.81 
tf.89 

12.08 
11.77 
11.41 
11.33 
10.69 
10.01 

250 

244 
237 
234 
221 

207 

Effect  of  Superheat  on  Steam  Engine  Economy. — Since  tur- 
bine tests  are  nearly  all  made  with  superheated  steam,  and  engine 
tests  almost  invariably  with  saturated  steam,  it  will  be  useful  to 


*See  paper  by  Prof.  Storm  Bull,  Journal  of  (he  Western  Sjctety  of  Engineer s^  De- 
.cember  1,  1903. 


196  STEAM  TURBINES 

have  some  means  for  estimating  the  steam  consumption  of  en- 
gines under  the  supposition  that  superheated  steam  is  used.  Data 
for  this  are  afforded  by  tests  upon  a  Belgian  engine  of  250  horse- 
power, which  has  established  a  remarkable  record  for  economy. 
The  tests  are  summarized  in  Table  XXL,  and  undoubtedly  are 
reliable,  as  they  were  made  by  Professor  Schroeter,  one  of  the 
most  experienced  experimenters  abroad.  The  tests  start  with 
saturated  steam  and  show  the  extremely  low  steam  consumption, 
for  an  engine  of  this  size,  of  12.08  pounds  per  indicated  horse- 
power hour.  The  items  following  this  one  give  the  consumption 
for  different  degrees  of  superheat.  The  results  show  that  for 
every  100  degrees  superheat  the  steam  consumption  per  horse- 
power per  hour  was  reduced  one  pound,  or  8.5  per  cent;  and 
that  the  consumption,  expressed  in  terms  of  equivalent  con- 
sumption of  saturated  steam,  was  reduced  %0  pound. 

Comparing  Turbine  and  Engine  Results. — The  reader  has  at 
his  disposal  in  the  last  three  tables,  together  with  Table  XVIIL, 
sufficient  information  to  form  an  opinion  upon  the  comparative 
rate  of  steam  consumption  of  turbines  and  engines  when  operat- 
ing at  their  most  economical  loads.  Due  allowances,  however, 
must  be  made  for  sizes  of  machines,  conditions  of  operation,  such 
as  steam  pressure,  vacuum,  superheat,  etc.  This  is  very  impor- 
tant, as  entirely  erroneous  opinions  are  often  formed  where  such 
allowances  are  not  made.  Taking  12.5  pounds,  previously  men- 
tioned, as  a  conservative  figure  for  the  most  economical  engines 
using  saturated  steam;  and  13.5  pounds  as  a  safe  figure  for  the 
average  high-grade  engine,  we  should  then  have  the  estimated 
rate  of  consumption  for  each,  with  different  degrees  of  super- 
heat, as  follows,  taking  the  Belgian  figures  as  a  basis : 

Most  Economical  Engine. 

With  Saturated        100  Degrees-       200  Degrees        300  Degrees 
Steam.  Superheat.  Superheat.  Superheat. 

12.5  11.4  10.4  9.3 

Average  Pligh-Grade  Engine. 
13.5  12,4  11.2  10.1 


STEAM  TURBINE  PERFORMANCE  197 

Comparing  these  figures  with  those  of  Table  XVIIL,  it  seems 
probable  that  the  reciprocating  engine  will,  under  exceptionally 
good  conditions,  show  a  little  better  economy  than  the  turbine, 
when  running  at  its  most  economical  load ;  but  that  what  we  have 
called  the  "average  high-grade  engine"  appears  to  about  equal 
the  turbine  in  its  rate  of  steam  consumption  at  most  economical 
loads.  In  the  next  chapter  the  question  of  variable  loads  will  be 
considered. 

Economy  of  Small  Engines  and  Turbines. — The  author  has 
several  times  seen  it  stated  by  English  engineers  that  in  sizes  of 
500  Kw.  and  less  the  engine  is  more  economical  than  the  turbine, 
but  that  as  sizes  increase  the  economy  of  the  turbine  improves 
more  rapidly  than  that  of  the  engine,  and  in  the  larger  powers  the 
turbine  is  equal  or  superior  to  the  engine.*  This  view  seems  to 
be  borne  out  by  the  facts  if  the  four-valve  compound  type  of  en- 
gine be  taken  for  comparison.  Engines  of  this  type  of  from  300 
to  500  Kw.  capacity  are  exceptionally  economical  and  are  rea- 
sonably so  in  still  smaller  sizes.  The  same  is  true  of  the  Willans 
engine,  used  so  extensively  in  England.  When  we  come  to  the 
single-valve,  high-speed  engine,  however,  which  is  in  such  general 
use  in  this  country,  there  is  no  doubt  that  its  rate  of  steam  con- 
sumption can  easily  be  improved  upon  by  the  turbine.  The  Gen- 
eral Electric  Company,  the  builders  of  the  Curtis  turbine,  recog- 
nize this  fact  and  do  not  provide  for  as  complete  expansion  of 
the  steam  in  turbines  of  small  sizes  as  in  their  larger  machines, 
since  it  is  not  necessary  to  do  so  in  order  to  compete  with  the 
high-speed  engine. 


*In  discussing  the  paper  by  Parsons,  Storey  and  Martin,  before  the  Institution  of 
Electrical  Engineers,  May,  1904,  E.  J.  Fox  said:  "Taking  figures  of  steam  economy 
as  given  by  Mr.  Parsons,  for  different  sizes  of  turbines,  there  is  no  difficulty  what- 
soever in  the  reciprocating  engine  giving  equally  good  results  up  to  the  1,500  Kw. 
size.  From  100  up  to  1,000  Kw.,  the  results  obtainable  with  reciprocating  engines  are 
better.  When  you  come  to  the  1,500  Kw.  size,  there  is  very  little  difference  between 
the  two;  and  finally  with  the  3,000  Kw.  size,  I  think  there  is  no  doubt  a  considerable 
difference  in  favor  of  the  turbine." 


CHAPTER  X 


STEAM  TURBINE  PERFORMANCE   (Continued). 
Characteristics  of  Turbines  Under  Variable  Loads. 

In  a  steam  engine,  and  in  certain  turbines,  like  the  Parsons,  in 
which  latter  there  is  an  auxiliary  valve  to  admit  steam  to  the  low- 
pressure  end  in  case  of  heavy  overloads,  the  lowest  steam  con- 


IUU1 

COO 
800 

100 
COO 
500 
400 
300 

c^ 

50* 

OVE 

RLC 

AD 

/ 

/ 

J 

25* 

OVE 

:RLC 

AD 

( 

MO 

5J_E 

c.ot 

OMI 

CAL 

-L0>! 

^D_ 

\ 

\ 

\ 

\ 

2b^ 

UNL 

hHL 

OAU 

N 

N 

^^^ 

50 

^  UNDER 

LO^ 

6> 

^<, 

8 

12  13  14  15  16 

LB.    STEAM  PER  I.H.P.  PER  HOUR. 

Fig.  1.     Corliss  Engine  Curve. 

sumption  per  horse-power  hour  occurs  at  or  near  the  normal  load. 
A  decrease  in  the  load  below  its  normal  point  causes  an  increase  in 
the  rate  of  consumption;  and  an  increase  in  the  load  above  the 
normal  point  produces  a  like  effect,  though  to  a  less  degree. 

In  Fig.  1  is  a  steam  consumption  curve  for  a  Corliss  engine. 
The  most  economical  load  is  at  point  A,  situated  on  the  curve  at 
the  extreme  left,  and  points  C  and  B,  above  and  below  A,  re- 
spectively, are  both  to  the  right  of  A.  These  two  extreme  points 
represent  the  consumption  at  50  per  cent  overload  and  underload, 


STEAM  TURBINE  PERFORMANCE  (Continued) 


199 


respectively,  and  it  is  evident  that  a  greater  range  of  loads  is 
possible  than  if,  for  example,  the  maximum  load  were  at  point  A, 
where  the  highest  economy  occurs,  and  the  consumption  at  other 
loads  were  represented  entirely  by  that  part  of  the  curve  extend- 
ing from  A  to  B. 

This  latter  condition  exists  in  any  turbine  in  which  the  speed 


2000 
1000 
1800 

iroo 

1GOO 
1500 
1400 
1300 

1100 

900 
800 
700 
600 
500 
400 

<W* 

/' 

/ 

/ 

/ 

A4-BY 

-PA{ 

s-v, 

MrV 

-    r\t 

ENS 

I 

\ 

\ 

\ 

\ 

\s 

\ 

V, 

^ 

^ 

vs 

^ 

v^ 

^ 

"XB 

18     19    20     21     22     23    24     25     26    27'     28 
LB.  STEAM  PER  Kw.    HOUR. 

Fig.  2.     Parsons  Turbine  Curve. 

is  regulated  entirely  by  throttling  the  steam,  as  examination  of  the 
diagrams  shortly  to  be  referred  to  will  make  evident. 

Curve  for  the  Parsons  Turbine. — In  Fig.  2  is  a  diagram  show- 
ing graphically  the  rate  of  steam  consumption  of  the  1,250  Kw. 
Westinghouse-Parsons  turbine,  tests  on  which  were  summarized 
in  Chapter  IX.  The  curve  in  this  diagram  resembles  that  in 
Fig.  1  for  the  Corliss  engine.  The  point  where  the  by-pass  opens 
is  clearly  defined,  and  it  is  the  action  of  this  by-pass  which  gives 
the  resemblance  in  this  curve  to  the  steam-engine  curve.  If  there 


200 


STEAM  TURBINES 


were  no  by-pass  the  turbine  would  not  be  able  to  carry  a  load 
above  1,500  Kw.,  and  we  should  have  a  curve  extending  from 
A  to  B,  simply,  like  the  lower  part  of  the  engine  curve.  The  gov- 
ernor in  this  type  of  turbine  is  virtually  a  throttling  governor  and 
the  rate  of  steam  consumption  gradually  decreases,  as  the  load 
increases,  until  the  by-pass  opens,  when  the  rate  increases,  since 
the  steam  which  enters  the  low-pressure  end  through  this  valve  is 
not  used  to  so  good  advantage. 

Curves  for  Turbines  of  the  Rateau  Type. — In  Fig.  3  is  a  steam 
rate  curve  for  a  Rateau  turbine,  plotted  from  Table  XII. ,  Chap- 
ter IX.  This  turbine,  also,  is  regulated  by  a  throttling  gov- 
ernor and  the  curve  resembles  the  sections  from  A  to  B  of  the 
Corliss  engine,  and  the  Parsons  curves,  Figs.  1  and  2.  The  most 


ELECTRICAL  HO3SE  POWER  ^ 

| 

1 

A^ 

MC 

ST  ECOI 

«JOM 

CAl 

LO 

\D 

\ 

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RATED 

NOF 

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VfMOS 

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5            16             17             13             10            20             2 

LB.  STEAM  PER  E.H.P.  HOUR. 
Fig.   3.     Rateau  Turbine    Curve. 

economical,  and  also  the  maximum,  load  carried  by  the  Rateau 
turbine  is  at  point  A,  at  the  upper  end  of  the  curve,  Fig.  3.  The 
normal  load  is  at  point  C,  and  the  fractional  loads  are  at  the  suc- 
cessive points  indicated.  Turbines  of  this  type  cannot  be  strictly 
said  to  have  overload  capacity,  and  their  normal  rated  load  must 
be  fixed  at  some  point  below  the  most  economical  load,  in  order 
to  give  the  machine  the  equivalent  of  overload  capacity.  Pro- 
fessor Rateau,  however,  has  proposed  the  use  of  a  by-pass  valve, 


STEAM  TURBINE  PERFORMANCE  (Continued) 


201 


in  which  case  the  performance  in  respect  to  variable  loads  would 
be  substantially  the  same  as  in  the  Parsons  type  fitted  with  this 
device. 

In  Fig.  4  are  three  curves  plotted  from  tests  in  Chapter  IX. 
on  the  400  Kw.  Westinghouse-Parsons  turbine,  given  in  Table 
XVI.,  the  Rateau  turbine,  Table  XII.,  and  the  Zoelly  turbine, 
Table  XIII.  The  curves  are  placed  in  their  correct  relative  posi- 
tions so  that  comparisons  can  be  made.  It  should  be  noted  that 


100 


14 


15 


1C  IT  18  19  20 

LB.  STEAM    PER  E.H.P.   HOUR. 

Fig.  4. 


the  Rateau  and  Zoelly  curves  are  based  on  the  electric  horse- 
power and  are  modified  somewhat  by  the  efficiency  of  the  gen- 
erator, while  the  Westinghouse  curve  is  based  on  brake  horse- 
power. The  similarity  of  the  curves,  however,  will  be  apparent. 
Curve  for  the  Curtis  Turbine. — This  turbine  operates  under 
boiler-pressure  steam  at  all  loads,  the  governor  changing  the  noz- 
zle area  instead  of  reducing  the  pressure  in  the  steam  chest,  as  in 
the  other  types  mentioned.  In  Fig.  5  is  a  curve  plotted  from  the 
Curtis  turbine  tests  of  Table  XIV.  The  least  steam  consumption 
occurs  at  the  normal  load  of  500  Kw.,  and  above  normal  load  the 
curve  bears  to  the  right  to  point  a,  just  as  in  the  steam-engine 
curve,  Fig.  1 ;  though  theoretically  it  should  continue  up  to  point 
b.  From  y^  load  to  l1/^  load  the  change  in  steam  consumption  is 


202 


STEAM  TURBINES 


very  slight,  but  from  ^  load  to  >4  load  it  is  rapid.  The  increased 
consumption  at  light  loads  in  this  turbine  is  due  to  internal  losses, 
such  as  friction,  diffusion  and  eddying  of  the  jets,  radiation,  etc., 
which  are  nearly  constant  and  therefore  absorb  a  larger  per- 
centage of  the  power  at  light  loads  than  at  heavy  loads.  Com- 
paring this  curve  with  that  of  the  Rateau  turbine,  we  find  the  lat- 
ter resembles  that  part  of  the  Curtis  curve  lying  between  a  point 
at  about  %  load  and  %  load.  The  upper  part  of  the  Curtis  curve, 
where  there  is  considerable  variation  in  power  with  only  slight 
variation  in  rate  of  steam  consumption,  is  absent  in  the  Rateau 


GOO 

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20  21  22  23  21  25 

STEAM   PER  K.W.  HOUrt,  LB. 

Fig.  5.     Curtis  Turbine  Curve. 

curve.  This  is  explained  by  the  fact  that  in  the  Rateau,  or  any 
turbine  which  governs  by  throttling  only,  there  are  not  only  the 
constant  internal  losses  just  mentioned  in  connection  with  the 
Curtis  turbine,  but  there  are  also  losses  due  to  the  throttling  of 
the  steam,  and  the  two  together  cause  the  steam  rate  to  increase 
more  rapidly  with  a  drop  in  load  than  in  the  Curtis  turbine,  where 
the  steam  is  not  throttled. 

Curve  for  the  De  Laval  Turbine. — In  the  De  Laval  turbine  con- 
ditions are  very  similar  to  those  found  in  the  Curtis  turbine. 
While  regulation  for  small  changes  in  load  is  effected  auto- 
matically by  throttling,  for  wide  variations  in  load  the  several 
steam  nozzles  are  opened  or  closed  by  hand,  as  required.  Two 


STEAM  TURBINE  PERFORMANCE  (Continued) 


203 


curves  are  plotted  in  Fig.  6,  one  for  the  last  series  of  tests  in 
Table  VII.  and  one  for  the  first  series  in  Table  VIII.  in  Chapter 
IX.  The  De  Laval  curves  have  the  characteristics  of  the  Curtis 
curve,  but  in  these  cases,  at  least,  the  De  Laval  turbine  will  evi- 
dently run  at  proportionately  lighter  loads  without  marked  in- 
crease in  steam  consumption. 


14  15  16  17 

X   LB.   PER  E.H.P.   HOUR. 
O  LB.   PER  B.H.P.   HOUR. 

Fig.  6.     De  Laval  Turbine  Curves. 

Results  of  Turbine  and  Engine  Tests  Under  Variable  Loads. 

Data  for  Performance  under  Variable  Loads. — Having 
studied  the  characteristics  of  different  types  of  turbines  under 
varying  loads,  it  is  now  in  order  to  institute  comparisons  be- 
tween reciprocating  engines  and  turbines  under  these  conditions. 
For  this  purpose  figures  will  be  given  from  the  results  of  a  num- 
ber of  tests  upon  engines  of  different  types.  The  tests  selected 
for  the  purpose  are  those  enumerated  below,  numbered  from  1  to 
6  for  convenience  in  reference. 

1.  A  slow-running,  simple  Corliss  engine  of  150  horse-power  at  Creusot. 
Steam  pressure,  60  pounds ;  vacuum,  27  inches ;  revolutions,  60.  Tests  re- 
ported in  Peabody's  Thermodynamics. 

2.  A  simple,  non-condensing  Corliss  engine  with  two  16-inch  cylinders, 
42   inches    stroke.     Pressure,    100   pounds;    revolutions,   86.      Quoted    in 
Engine  Tests  by  Geo.  H.  Barrus. 

3.  A  tandem-compound,  condensing  Buckeye   engine,   160  horse-power. 


204  STEAM  TURBINES 

Cylinders,  11  and  19  by  24;  revolutions,  160.    Tests  given  by  A.  K.  Mans- 
field, Proc.  A.  S.  M.  R,  1897. 

4.  A  McEwen,  tandem-compound,  non-condensing,  high-speed  engine. 
Cylinders,  9  and  16  by  14;  pressure,  112  pounds;  revolutions,  265.    Tested 
by  Prof.  R.  C.  Carpenter,  Proc.  A.  S.  M.  R,  1893. 

5.  A    Fleming,    four-valve,,  tandem-compound,    condensing    engine,    of 
500  horse-power.     Cylinder  ratio,  1  to  7.33;  pressure,  150  pounds.     Re- 
ported by  B.  T.  Allen,  Proc.  A.  S.  M.  R,  1904.     These  tests  have  been 
criticised  because  the  steam  pressure  was  allowed  to  drop  in  the  boiler  at 
light  loads.     It  is  probable,  however,  that  this  variation  did  not  greatly 
affect  the  results,  since  the  high-pressure  admission  valves  throttled  the 
steam  at  light  loads,  so  that  full  boiler  pressure  could  not  have  been  real- 
ized in  the  cylinder  in  any  case. 

6.  A  Rice  and  Sargent,  compound  engine  (Corliss  type)  with  cylinders 
20  and  40  by  42  inches.     Steam  pressure,  150  pounds;  vacuum,  28  inches; 
revolutions,  120.    Tested  by  Prof.  D.  S.  Jacobus  for  the  builders. 

In  obtaining  figures  from  the  tests  upon  the  above  engines,  and 
also  from  the  turbine  tests  in  this  chapter,  the  method  followed 
has  been  to  plot  curves  for  the  rate  of  steam  consumption  under 
varying  loads,  and  from  those  curves  to  take  the  figures  at  such 
points  as  were  desired.  For  example,  the  results  of  an  engine 
test  might  not  give  the  steam  consumption  at  exactly  half  load, 
but  by  plotting  the  curve  for  such  results  as  were  given,  the  half- 
load  consumption  could  be  approximately  determined.  Figures 
will  be  given  without  showing  the  curves,  except  where  they  have 
previously  been  printed.* 


•The  figures  given  below  are   the  tabulated  results  from  which   the  curves   for   the 

engine  tests  were  plotted.  Of  each  group  of  figures,  the  first  column  contains  the 
indicated  horse-powers  and  the  second  column  the  corresponding  steam  consumptions 
in  pounds  per  horse-power  hour. 

Engine  No.  i.                          Engine  No  2.  Engine  No.  3. 

194           18.6                           342.4         25.91  200.5         17.55 

175           17.7                           287.1         25.39  172.           16.3 

150           17.3                            222.2         25.83  153.7         17.1 

117           17.6                           146.2         31.43  140.9         17.65 

91.7        18.5                           100.4         38.38  135.9         17.76 

37.           73.63  103.           18.2 
89.5         18.25 
55.1         24.1 

Engine  No.  4.                          Engine  No.  5.  Engine  No.  6. 

126.          21.2                           553.49       12.73  1,004.3         12.75 

118.4        20.3                           501.55       12.66  853.3         12.33 

100.6        21.7                           348.28       12.33  819.6         12.55 

95.1  19  1                           321.54       13.59  627.4         12.10 
80.5        18.9                             87.07       14.42  491.4         13.92 
61.          19.69  339.7         14.58 
44.          20.7 

27.2  23.11 


STEAM  TURBINE  PERFORMANCE  (Continued)  205 

The  turbine  tests  selected  for  comparison  are  the  following, 
taken  from  data  in  Chapter  IX. : — 

No.  1.  300  H.  P.  De  Laval.  Fourth  test  in  Table  VII.,  with  superheated 
steam.  Results  in  terms  of  electrical  horse-power. 

No.  2.  300  H.  P.  De  Laval.  First  test  in  Table  VIII.,  with  saturated 
steam.  Results  in  terms  of  brake  horse-power. 

No.  3.  500  H.  P.  Rateau.  Table  XII.  Results  in  electrical  horse- 
power units. 

No.  4.  500  H.  P.  Zoelly.  Table  XIII.  The  results  in  the  table  were  con- 
verted into  terms  of  electrical  horse-power  before  plotting  curve. 

No.  5.  500  Kw.  Curtis.  Table  XIV.  Results  in  pounds  per  kilowatt 
hour. 

No.  6.  500  KW.  Curtis  turbine  at  Newport,  R.  I.  The  figures  for  this 
machine  were  taken  from  a  curve  plotted  from  the  results  of  tests  by 
Geo.  H.  Barrus  and  given  in  a  paper  by  W.  L.  R.  Emmet  before  the 
Engineers'  Club  of  Philadelphia,  in  March,  1904. 

No.  7.  1,250  Kw.  Westinghouse-Parsons,  with  by-pass  valve.  Table 
XVII.  Results  in  terms  of  kilowatts. 

No.  8.  400  Kw.  Westinghouse-Parsons.  First  and  third  tests,  Table 
XVI.,  one  with  superheated  steam  and  one  with  saturated  steam.  Re- 
sults in  terms  of  brake  horse-power.  This  turbine  had  no  by-pass. 

Comparison  of  Tests  under  Variable  Loads. — One  way  of  com- 
paring tests  under  variable  loads  is  to  set  a  percentage  limit  for 
the  rate  of  steam  consumption  and  then  determine  how  great  a 
variation  in  load  the  engine  or  turbine  will  permit  without  ex- 
ceeding this  limit. 

The  author  has  assumed  a  limit  of  10  per  cent  increase  in  the 
rate  of  steam  consumption  above  the  most  economical  rate,  and 
then  determined  the  approximate  variation  in  power  for  each 
turbine  or  engine,  corresponding  to  the  10  per  cent  variation  in 
the  steam  rate. 

The  variation  in  power  was  found  in  per  cent  of  the  maximum 
power  developed  by  each  turbine  or  engine,  and  is  as  follows : 
Turbine  No.  1.     Variation  in  power,  55%  of  maximum  power. 

No.  2.  "  "       "        60%    " 

No.  3.  "  "       "        35%    " 

No.  4.  "  "       "        35%    " 

No.  5.  "  "       "        60%    " 

No.  6.  "  "       "        60%    " 

No.  7.          '"  "       "        55%    " 

No.  8.  "  "       "        50%    " 


206  STEAM  TURBINES 

On  the  same  basis  we  have  the  following  results  from  the  en- 
gine tests  of  which  a  list  has  been  given : 

Engine    No.  1.  Variation  in  power,  60%  of  maximum  power. 
No.  2.  "  "       "        45%    " 

No.  3.  '    "       "        50%    " 

No.  4.  "  "       "        65%    " 

No.  5.  "       "        70%    " 

No.  6.  "       "        55%    " 

In  reviewing  these  results,  it  seems  fair  to  discard  the  figures 
for  turbine  tests  Nos.  3,  4  and  8,  since  these  particular  turbines 
were  not  fitted  with  a  by-pass  to  admit  high-pressure  steam  to  the 
low-pressure  end,  but  are  of  a  type  requiring  the  by-pass  to  give 
the  best  results.  The  remaining  tests  indicate  that  for  a  10  per 
cent  variation  in  the  rate  of  steam  consumption  the  variation  in 
power  developed  by  the  turbines  averages  55  to  60  per  cent. 

The  engine  tests  quoted  give  an  unusually  good  presentation  of 
the  claims  of  the  steam  engine.  The  six  tests  are  from  five  differ- 
ent types  of  engines,  and  include  both  simple  and  compound. 
The  only  result  apparently  open  to  question  is  the  one  of  70  per 
cent  variation  for  engine  No.  5.  Only  one  test  was  made  upon 
this  engine  at  very  light  loads,  and  as  the  only  way  of  checking 
that  test  is  by  means  of  the  curve  drawn  through  points  plotted 
for  considerably  larger  loads,  the  result  of  70  per  cent  does  not 
appear  well  established.  This  engine  ought  to  do  as  well  as  en- 
gine No.  4,  however,  wrhich  gave  65  per  cent.  If  we  call  the 
variation  65  per  cent,  instead  of  70  per  cent,  the  average  engine 
result  for  the  whole  group  will  lie  between  55  and  60  per  cent,  or 
the  same  as  in  the  case  of  the  turbines. 

Direct  Comparison  of  Engine  and  Turbines. — In  Fig.  7  are 
plotted  three  curves  showing  the  rate  of  steam  consumption,  un- 
der different  loads,  of  a  Curtis  turbine,  a  Westinghouse-Parsons 
turbine,  and  a  Rice  and  Sargent  engine.  The  first  is  the  500  Kw. 
turbine  at  Newport,  R.  I.,  tested  by  George  H.  Barrus;  the 
second  the  1,250  Kw.  turbine  reported  in  Table  XVII.,  and  the 
third  the  850  H.  P.  compound  engine  referred  to  previously  in 


STEAM  TURBINE  PERFORMANCE  (Continued) 


20; 


Chapter  IX.*  Fortunately  the  tests  upon  this  engine  included  a 
record  of  the  electrical  output,  from  which  the  rate  of  consump- 
tion per  electrical  horse-power  hour  could  be  calculated,  enabling 
direct  comparisons  to  be  made  with  turbines  without  having  to 
allow  for  the  efficiency  of  the  apparatus. 


13  14  15  16  17  18  19  20 

Pounds  Steam  per  Elertrifal  Horse  Power  per  Hour 

Fig.  7. 


*The  curve  for  the  Curtis  turbine  was  plotted  from  a  curve  showing  the  results  of 
the  Barrus  tests,  given  by  W.  L.  R.  Emmet  of  the  General  Electric  Company,  in  a 
paper  presented  before  the  Engineers'  Club  of  Philadelphia,  March,  1904.  The  curve 
for  the  Rice  and  Sargent  engine  was  plotted  from  a  report  of  Professor  Jacobus'  tests 
upon  this  engine,  issued  by  the  builders  in  pamphlet  form.  The  figures  are  as  follows: 


Electrical 
Horse  Power. 
945.6 
796.4 
761.1 
564.4 
440.2 
282.9 


Total  Weight 

Steam  per  Hour. 

12,803 

10,520  „ 

10,283 

7,591 

6,840 

4,953 


Pounds  Steam 
per  E.  H.  P.  per  Hour. 
13.54 
13.21 
13.51 
13.45 
15.54 
17.51 


208  STEAM  TURBINES 

Deductions  from  Curves  of  Fig.  7. — The  Curtis  turbine  pro- 
duces the  flattest  curve,  while  there  is  not  much  to  choose  between 
the  Rice  and  Sargent  and  the  Westinghouse  curves,  so  far  as  this 
feature  is  concerned.  In  point  of  economy,  the  engine  is  easily  in 
the  lead.  This  engine,  however,  has  proven  itself  to  be  excep- 
tionally economical,  while  the  500  Kw.  Curtis  turbine  as  then 
constructed  had  only  two  stages  and  would  not  run  with  as  low 
steam  consumption  as  the  larger,  three-  or  four-stage  machines. 
It  would  lead  to  erroneous  conclusions  to  accept  this  diagram  too 
literally  as  an  example  of  what  may  be  expected  from  engines  and 
turbines  in  general,  when  operating  with  variable  loads.  The  dia- 
gram is  a  significant  one,  however,  and  inevitably  leads  to  the 
conclusion  that  a  compound  Corliss  engine  is  able  to  hold  its  own, 
in  comparison  with  turbines,  provided  the  variation  in  load  is  not 
over  40  or  50  per  cent  above  or  below  normal  load. 

Engine  Performance  from  an  Operative  Standpoint. — In  con- 
nection with  the  above  comparisons  of  turbines  and  engines  under 
variable  loads,  the  author  wishes  to  point  out  two  important 
facts :  First,  that  the  water  rate  load  curve  of  an  engine  does  not 
represent  its  true  performance  where  the  load  is  rapidly  fluc- 
tuating. The  rapidly  changing  cylinder  conditions  that  exist 
when  the  load  is  fluctuating  lead  to  increased  condensation  and 
consequent  waste  of  steam  that  does  not  occur  when  an  engine  is 
under  test  at  certain  fixed  loads. 

Second,  that  in  the  tests  quoted  the  overloads  were  not  more 
than  50  per  cent  above  normal  load  and  in  most  cases  were  less 
than  this.  In  one  way  this  adds  to  the  interest  of  the  comparisons 
because  it  shows  what  turbines  will  do  within  ranges  of  load  under 
which  reciprocating  engines  generally  operate ;  but  in  another  way 
it  is  not  fair  to  the  turbine,  because  the  practice  now  is  to  design 
turbines  to  carry  much  heavier  overloads  than  can  the  steam  en- 
gine. It  is  not  usual  for  engines  to  run  with  loads  greater  than 
50  per  cent  in  excess  of  their  normal  load,  it  being  advisable  to 
put  in  a  larger  engine  when  this  point  is  reached.  Corliss  engines 
fitted  with  two  eccentrics  can  carry  overloads  as  great  as  100 
per  cent,  provided  the  vacuum  can  be  maintained,  but  not  with 
good  economy.  The  tendency  always  is  to  install  a  steam 
engine  large  enough  to  safely  handle  the  overloads,  and  then  let  it 


STEAM   TURBINE  PERFORMANCE    (Continued)  209 


operate  at  an  average  load  considerably  below  the  normal  load  for 
the  balance  of  the  time.  This  is  illustrated  in  a  striking  manner 
by  tests  made  by  students  of  Cornell  University  upon  35  street 
railroad  power  plants  during  a  period  of  12  years.  These  results 
have  been  gathered  by  Prof.  R.  C.  Carpenter  and  grouped  accord- 
ing to  the  type  of  engine.*  There  were  eight  tests  of  compound 
condensing  engines  of  the  Corliss  and  similar  types,  and  the  main 
results  of  these  are  given  in  Table  I.  They  show  the  same  char- 
acteristics as  the  others  of  the  35  tests,  not  quoted  here. 

TABLE  I. 

SUMMARY  OF  TESTS  ON  COMPOUND  CONDENSING  ENGINES  OF  THE  CORLISS 
AND  SIMILAR  TYPES,  IN  STREET  RAILWAY  PLANTS. 


H.  P.  of 

Engine. 

Steam  per 
1  H.  P. 
per  Hour. 

Mean 
Observed 
H.  P. 

Per  cent  of 
Mean  H.  P. 
to  Capacity. 

825 

1,000 
1,000 
350 
500 
2,000 
200 
1,600 

22.7 
21.9 
20.0 
16.64 
16.90 
14.50 
17.30 
20.50 

482 
277 
314 
182 
290 
814 
145 

58.2 
27.7 
31.4 
52.2 
58. 
40.7 
72. 

Average 

18.80 

48.6 

Taking  the  average  results,  the  average  load  on  these  engines 
was  less  than  half  their  normal  load,  and  they  were  running  day 
in  and  day  out  on  a  steam  consumption  of  18.80  pounds  per  horse- 
power hour,  instead  of  the  13  or  14  pounds  that  similar  engines  are 
capable  of  at  normal  load,  if  in  good  condition. 

How  the  Turbine  Improves  upon  Engine  Performance. — What 
is  the  answer  of  the  turbine  to  this  condition  of  affairs?  The 
turbine,  first  of  all,  is  not  subjected  to  serious  losses  from  internal 
condensation  and  under  a  rapidly  fluctuating  load  should  show 
nearly  the  same  economy  as  indicated  by  the  water  rate  load 
curve.  This  is  of  great  importance  in  street  railway  work. 

Second,  the  results  that  can  be  secured  with  a  turbine  under 
heavy  overloads  are  indicated  by  a  test  upon  a  Westinghouse- 


*Sibley  Journal  of  Engineering,   December,   1904. 


210 


STEAM  TURBINES 


Parsons  400  Kw.  turbine,  by  F.  P.  Sheldon  &  Co.,  mechanical 
engineers,  Providence,  R.  I. 

This  turbine  showed  the  usual  results  under  loads  varying  from 
YZ  to  V/2  the  normal  rating;  but  in  addition — and  this  is  the  im- 
portant point — demonstrated  its  ability  to  carry  an  overload  of 
100  per  cent  with  an  increase  in  the  rate  of  steam  consumption  of 
less  than  10  per  cent. 

TABLE   II. 

PERFORMANCE  OF  WESTINGHOUSE-PARSOXS  TURBINE  UNDER  VARIABLE 
LOADS,  INCLUDING  HEAVY  OVERLOADS. 

Steam  Pressure  i^o  Lb.  Absolute;   Vacuum  28  Inches. 


Load  on 
Turbine. 

Steam  Consumption  —  Pounds 
per  B.  H.  P.  per  Hour. 

Saturated  Steam. 

Steam  Superheat- 
ed 100  Degrees. 

V2  Load  
%  Load  
Full  Load  
1#  Load  
\Yz  Load 

15.86 
15.05 
13.89 
13.85 

isiia 

14.34 
13.45 
12.48 

12.41 
12.79 
13.55 

100$  Overload  

The  meaning  of  this  is  that  a  turbine  can  be  installed  with 
reference  to  the  average  load  that  it  has  to  carry,  instead  of 
with  reference  to  the  maximum  load,  as  in  the  case  of  the  steam 
engine,  and  then  trust  to  the  by-pass  to  take  care  of  the  overloads. 
The  turbine  will  then  be  operating  at  or  near  its  point  of  best 
economy  most  of  the  time,  and  being  of  smaller  power  than  a 
steam  engine  for  the  same  work,  will  not,  at  light  loads,  drop  to  so 
small  a  percentage  of  the  normal  load  as  will  the  steam  engine. 
The  recent  practice  of  encasing  the  turbine  generator  and  cooling 
it  with  forced  air  circulation  makes  the  generator  amply  able  to 
handle  overloads  and  the  generous  condensing  systems  used  with 
turbines  should  preclude  a  serious  drop  in  vacuum  under  excessive 
overloads. 

Taking  into  account  the  operative  conditions  of  engines  and 
turbines  as  they  exist,  the  author  is  inclined  to  the  opinion  that 
the  turbine  will  surpass  the  engine  under  fluctuating  loads.  This 
conclusion  applies  directly  to  the  Parsons  type,  fitted  with  a  by- 
pass valve.  Enough  data  have  not  been  published  in  respect  to 


STEAM   TURBINE  PERFORMANCE    (Continued)  211 

the  Curtis  and  other  types  to  show  how  the  matter  stands  with 
them,  although  the  guarantees  made  by  the  builders  of  the  Curtis 
turbine  are  all  that  could  be  desired. 


The  Effect  of  Vacuum  Upon  Economy. 

In  the  operation  of  a  steam  engine,  it  is  customary  to  run  with  a 
vacuum  of  about  26  inches  in  the  condenser.  While  a  higher 
vacuum  is  slightly  advantageous  from  the  fact  that  it  reduces  the 
back  pressure  upon  the  piston,  the  reciprocating  engine  is  not 
capable  of  taking  full  advantage  of  the  increased  vacuum,  as  ex- 
plained in  the  chapter  upon  condensers.  In  a  turbine,  however, 
steam  can  easily  be  expanded  to  the  volume  corresponding  to  the 
lowest  pressure  attainable  in  a  condenser,  and  the  effort  is  made  to 
run  with  a  vacuum  of  28  or  29  inches. 

We  are  now  concerned  with  the  gain  in  economy  resulting  from 
this  higher  vacuum,  or  what  is  more  to  the  point,  with  the  loss  in 
economy  due  to  a  reduction  in  the  vacuum  below  the  usual  figure 
of  28  inches  at  which  turbines  are  designed  to  run.  While  nearly 
all  turbine  tests  are  made  with  a  vacuum  of  28  or  29  inches,  and 

TABLE  III. 
ECONOMY  OF  750  Kw.  WESTINGHOUSE-PARSONS  TURBINE  WITH  26-lNCH 

AND    28-lNCH    VACUUM. 


Approximate  Load 
in  E.  H.  P. 

Pounds  Steam  per  E.  H.  P.  per  Hour. 

Increase  in  Steam 
Consumption  Due  to 
Low  Vacuum. 

With  28-Inch 
Vacuum. 

With  26-Inch 
Vacuum. 

Tests  with  Sa 

turated  Steam. 

1,500 
1,100 
775 
500 

13.90 
13.76 
14.65 
15.95 

14.73 
15.06 
16.64 
18.35 

5.  3  per  cent 
9.5  per  cent 
13.  4  per  cent 
15.0  per  cent 

T 

ests  with  Steam  Su 

per  heated  130  Degre 

es. 

1,500 
1,2^5 
1(25 
525 

11.50 
11.42 
11.79 
13.85 

12.98 
13.05 
13.18 
16.04 

12.9  per  cent 
14.  3  per  cent 
11.  8  per  cent 
15.  8  per  cent 

Average  Results  with 
Saturated  Steam... 
Average  Results  with 
Superheated  Steam 

14.59 
13.81 

16.19 
12.14 

10.9  per  cent 
13.7  per  cent 

212 


STEAM  TURBINES 


figures  are  usually  quoted  on  this  basis,  it  is  not  always,  and  we 
doubt  if  it  is  usually,  possible  to  maintain  so  good  a  vacuum  in 
commercial  operation.  Accordingly,  what  we  want  to  know  is, 
how  much  the  steam  consumption  of  a  turbine  will  be  increased  by 
this  drop  in  vacuum. 

Turbine  Tests  with  Different  Vacuums. — The  most  complete 
tests  of  turbines  under  different  vacuums  have  been  made  at  the 
shops  of  the  Westinghouse  Machine  Company,  and  are  reported  in 
their  turbine  catalogue.*  In  Table  III.  is  a  summary  of  tests 
upon  a  750  Kw.  turbine  compiled  from  this  source,  and  in  Table 
IV.  is  a  summary  of  tests  upon  a  1,250  Kw.  unit. 

Table  III.  gives  results  under  different  loads  for  vacuums  of 
28  and  26  inches,  respectively,  and  with  both  saturated  and  super- 
heated steam.  It  is  evident  that  the  increase  in  steam  consumption 
is  more  marked  at  light  than  at  heavy  loads,  and  the  average  in- 
crease for  the  two  inches  difference  of  vacuum  is  10.9  per  cent 
with  saturated  steam  and  13.7  per  cent  with  superheated  steam. 

TABLE  IV. 

ECONOMY  OF  1,250  Kw.  WESTINGHOUSE-PARSONS  TURBINE  WITH  DIFFERENT 
VACUUMS.     TESTS  WITH  SATURATED  STEAM. 


Pounds  Steam  per  E.  H.  P.  per  Hour. 


Approximate 
Load  in  E.  H.  P. 

With  28-Inch 
Vacuum. 

With  27-Inch 
Vacuum. 

With  26-Inch 
Vacuum. 

With  25-Inch 
Vacuum. 

2,rOO 
1,200 
400 

14.73 
15.^5 
21.8 

15.22 
17.06 
24.33 

15.72 
18.29 
26.62 

16.32 
18.82 
28.1 

Average  Results.  .  . 
Differences  

17.43 

18.88 
1.45 

20.21 
1.33 

21.08 
0.87 

Per   cent   Increase 
for  Each  inch 
Drop  in  Vacuum 



8.3 

7.1 

4.3 

Increase  in   Steam  Consumption  on  account  of  Drop  in   Vacuum  from  28  to 
inches: 

Load,  2,000  E.  H.  P.,  Increase  =  6.7  per  cent. 
Load,  1,200  E.  H.  P.,  Increase  =  16.1  per  cent. 
Load,  400  E.  H.  P.,  Increase  =  22.1  per  cent. 
Average  Increase  =  15.9  per  cent. 


*In  a  turbine,  the  benefit  derived  from  a  good  vacuum  is  much  more  than  in  a 
reciprocating  engine,  every  one  inch  of  vacuum  between  23  inches  and  28  inches 
affecting  the  consumption  on  an  average  about  3  per  cent  in  a  100  Kw.,  4  per  cent  in 
a  500-  Kw.,  and  5  per  cent  in  a  1,500  Kw.  turbine,  the  effect  being  more  at  high 
vacuum  and  less  at  low. — Hon.  Chas.  A.  Parsons  in  a  paper  before  the  Institution  of 
Electrical  Engineers,  May,  1904. 


STEAM  TURBINE  PERFORMANCE  (Continued) 


213 


Table  IV.  gives  results  under  different  loads  for  vacuums 
varying  by  inches  from  28  to  25  inches.  The  average  increase  in 
steam  consumption  when  the  vacuum  drops  from  28  to  27  inches 
is  8.3  per  cent;  from  27  to  26  inches,  7.1  per  cent;  from  26  to  25 
inches,  4.3  per  cent,  which  shows  that  the  drop  from  28  to  27 
inches  affects  the  results  nearly  twice  as  much  as  the  drop  from 
26  to  25  inches.  The  average  increase,  as  summarized  at  the  bot- 
tom of  the  table,  for  a  drop  in  vacuum  from  28  to  26  inches,  is 
15.9  per  cent.  We  see  from  this,  and  the  previous  table  that  a  loss 
of  two  inches  of  vacuum,  below  28  inches,  causes  in  these  cases 
an  increase  in  steam  consumption  of  say  from  10  to  15  per  cent. 
A  turbine,  therefore,  running  on  20  pounds  of  steam  per  Kw. 
hour,  with  28  inches  vacuum,  might  be  expected  to  use  from  22  to 
23  pounds  if  the  vacuum  was  reduced  to  26  inches — quite  a  likely 
condition. 

The  diagram,  Fig.  8,  was  plotted  from  the  tests  upon  the  1,250 


Vacuum  in  Inches. 
Fig.  8. 


214     .  STEAM  TURBINES 

H.  P.  turbine  and  shows  the  steam  consumption  at  different 
vacuums.  Curve  A  is  for  the  load  of  400  H.  P. ;  curve  B  foi 
1,200  H.  P. ;  curve  C  for  2,000  H.  P. ;  and  the  dotted  curve  D  for 
the  average  of  the  three  loads.  This  chart  shows  graphically  the 
two  facts  already  pointed-  out,  that  a  change  in  vacuum  has  a 
greater  effect  at  light  than  at  heavy  loads,  and  that  the  effect  is 
more  marked  at  high  than  at  low  vacuums. 

The  Effect  of  Superheating. 

In  the  tests  upon  the  De  Laval  and  Parsons  turbines  in  the 
last  chapter  are  data  upon  the  gain  (at  the  turbine)  through 
the  use  of  superheated  steam.  While  such  results  have  little  or 
no  significance,  unless  computed  on  the  heat-unit  basis,  it  is  be- 
lieved to  be,  and  probably  is,  the  case,  that  superheated  steam  is 
desirable  for  turbine  use.  In  the  summary  of  tests  upon  a  300- 
horse-power  De  Laval  turbine  in  Table  IX.,  the  gain  from  super- 
heating is  about  1  per  cent  for  every  eight  to  10  degrees  super- 
heat. In  their  report  upon  the  tests  of  a  400  Kw.  Westinghouse- 
Parsons  turbine,  already  referred  to,  Dean  and  Main  estimate  the 
gain  from  superheating  to  be  one  per  cent  for  each  10  degrees,  up 
to  180  degrees  superheat,  which  was  the  limit  of  their  tests.  This 
figure  is  corroborated  by  Parsons,  who,  in  the  paper  before  the 
British  Institute  of  Electrical  Engineers,  from  which  Tables  X. 
and  XL  are  quoted,  says,  "Every  10  degrees  superheat  up  to  about 
150  degrees  F.,  affects  the  consumption  about  1  per  cent." 

Reason  for  Gain  from  Superheated  Steam. — The  gain  from 
superheated  steam  in  the  turbine  comes  from  reduced  friction  be- 
tween the  rapidly  flowing  steam  and  the  passages  of  the  turbine, 
and  between  the  steam  and  the  surfaces  of  the  rotating  wheels. 
In  the  steam  engine  the  gain  from  superheated  steam  is  due  to  the 
reduction  of  cylinder  condensation,  resulting  in  less  loss  from  re- 
evaporation  of  this  moisture  at  the  lower  pressures  during  the  lat- 
ter part  of  the  stroke,  and  during  exhaust.  The  initial  condensa- 
tion of  saturated  steam  entering  an  engine  cylinder  is  often  a? 
much  as  40  or  50  per  cent,  and  this  is  partly  or  entirely  prevented 
when  the  steam  is  superheated,  depending  upon  the  degree  or 
superheat.  In  the  turbine  the  effect  of  superheated  steam  is  als- 
to  reduce  condensation,  which  always  occurs  when  saturated 


STEAM  TURBINE  PERFORMANCE  (Continued)  215 

steam  expands  adiabatically,  as  it  is  assumed  to  expand  when  it 
flows  through  a  turbine.  But  the  walls  of  the  turbine  remain  at 
so  uniform  a  temperature  that  there  is  no  opportunity  for  loss 
through  reevaporation  of  this  condensed  steam,  and  the  gain  real- 
ized must  therefore  result  from  the  reduced  friction,  as  stated 
above. 

It  is  fallacious  to  suppose  there  is  any  considerable  gain  because 
superheated  steam  has  a  higher  temperature  than  saturated  steam. 
The  theoretical  or  thermodynamic  gain  from  superheating  is 
almost  nothing.  It  is  true  that  the  heat  represented  by  the  super- 
heat is  received  at  a  higher  mean  temperature  than  the  other  por- 
tions of  the  heat  in  the  steam,  and  there  is  a  theoretical  advantage 
from  this  fact.  The  amount  of  heat  received  at  this  high  tempera- 
ture, however,  is  so  small,  compared  with  the  amount  received 
before  superheating  begins,  as  to  be  almost  negligible.  One  pound 
of  saturated  steam  at  100  pounds  gauge  pressure  contains  nearly 
900  latent  heat  units ;  and  if  this  steam  were  then  superheated  1 00 
degrees,  it  would  acquire  only  about  six  heat  units  in  addition,  in 
spite  of  its  high  temperature. 

Practical  Considerations. — In  the  account  published  in  The 
Iron  Age,  May,  1904,  of  Barms'  tests  upon  one  of  the  Curtis 
turbines  at  Newport,  R.  L,  the  reported  results  with  superheated 
steam  were  not  entirely  favorable.  This  station  was  equipped  with 
a  separately  fired  Schmidt  superheater  and  unless  the  superheater 
was  worked  to  nearly  its  full  capacity  there  was  a  loss  from  its 
use.  This  is  not  to  condemn  a  moderate  degree  of  superheating, 
however,  such  as  can  readily  be  realized  by  the  use  of  superheating 
tubes  in  the  same  setting  with  a  steam  boiler,  and  such,  moreover, 
as  can  be  maintained  from  day  to  day  in  actual  operation.  The 
author  would  recommend  a  turbine  plant  to  be  so  equipped  with 
capacity  for  superheating  50  degrees.  This  ensures  perfectly  dry 
steam  at  the  turbine,  which  is  the  main  point  to  be  attained,  with- 
out trouble  from  high  temperatures. 

Economy  with  Change  in  Speed. 

The  turbine  is  a  one-speed  machine.  Its  speed  cannot  be 
changed  from  that  at  which  it  is  designed  to  run  without  serious 
loss  in  efficiency.  The  speed  of  the  rotating  buckets  must  bear  a 


216  STEAM  TURBINES 

definite  relation  to  the  velocity  of  the  steam,  as  determined  by  the. 
proportions  of  the  turbine,  or  otherwise  there  will  be  disastrous 
impact  of  the  steam  against  the  vanes.  In  the  Curtis  turbine  at 
the  Newport  station,  tests  were  run  to  determine  the  variation  in 
steam  consumption  with  a.  variation  in  the  speed.  At  1,900  revolu- 
tions per  minute,  the  water  rate  per  kilowatt  hour  was  about 
19.75  pounds;  at  1,600  revolutions,  about  20.75  pounds;  at  1,300 
revolutions,  about  22.7 ;  and  at  1,000  revolutions,  about  27  pounds. 
This  shows  that  at  first  the  decrease  in  speed  produced  only  a 
small  effect  on  the  economy;  but  later,  as  the  slower  speeds  are 
reached,  the  steam  consumption  increases  very  rapidly. 

In  a  paper  upon  steam  turbines*  by  Ernest  N.  Jansen,  tests  are 
quoted  upon  several  turbines  running  with  varying  speeds.  With 
a  400  Kw.  Westinghouse- Parsons  turbine  at  1,800  revolutions  per 
minute,  or  half  of  the  designed  speed,  there  was  an  increase  of 
25  per  cent  in  the  steam  consumption. 

•Published  in  the  Journal  of  the  Society  of  Naval  Architects,   1904. 


CHAPTER  XI 
EXPERIMENTS  ON  THE  FLOW  OF  STEAM. 

Napier's  Rules  for  the  Floiv  of  Steam. — R.  D.  Napier  was  one 
of  the  first  experimenters  to  secure  results  of  value  on  the  flow  of 
steam.  He  published  data  in  1866  to  show  that  when  steam  flows 
through  a  cylindrical  nozzle  having  a  rounded  inlet  the  weight  dis- 
charged in  a  given  time  depends  only  on  the  initial  absolute  pres- 
sure, so  long  as  the  absolute  pressure  against  which  the  nozzle  dis- 
charges does  not  exceed  %0  of  that  pressure.  Thus,  if  steam 
flows  from  a  pressure  of  100  pounds  absolute  to  a  pressure  of  60 
pounds  absolute,  the  weight  discharged  in  a  given  time  will  be 
practically  the  same  as  though  the  nozzle  were  discharging  at  some 
lower  pressure,  as  into  the  atmosphere,  or  into  a  partial  vacuum. 

If,  however,  the  final  pressure  is  more  than  60  pounds,  the 
weight  discharged  will  be  less  than  before  and  will  become  very 
much  less  as  the  difference  of  pressures  decreases. 

In  the  London  Engineer  for  November  26  and  December  3, 
1869,  Professor  Rankine  reviews  Napier's  work  and  presents  one 
of  the  best  theoretical  discussions  of  the  flow  of  steam  that  has  been 
published.  He  concludes  that  the  formulas  given  below,  commonly 
known  as  "Napier's  Rules,"  will  give  a  rough  approximation  of  the 
weight  of  steam  discharged  through  a  conoidal  converging  nozzle. 
The  rules  also  apply  in  the  case  of  short,  cylindrical  tubes  with 
rounded  inlets. 

Let  H^flow  in  pounds  per  second. 

/j^higher    pressure    and    />,=lower    pressure,    both    in 
pounds  per  square  inch  absolute,  and 

a=area  of  orifice  in  square  inches. 

Case  I. — Lower  absolute  pressure  equal  to  or  less  than  %0  of 
higher  absolute  pressure : 

p,  a 

W=-  (1) 

70 


218  STEAM  TURBINES 

Case  II.  —  Lower  absolute  pressure  more  than  %0  of  higher 
absolute  pressure  : 


Example.—  />1=100  ;  o=l. 

Case  I. 

100X1 

W=—      —=1.428  pounds  per  second. 
70 

Case  II.     Let  p2—SQ.    Then, 


tf/=( 80-^42)  XV  (100— 80) -f- 1  X80 
=1.9XV-375 
=1.16  pounds  per  second. 

Napier's  rules  give  better  results  in  cases  where  applicable  than 
the  more  complicated  rules  based  on  the  laws  of  thermodynamics. 

Brownlee's  Safety-Valve  Experiments. — The  next  important 
tests  to  be  recorded  are  described  in  a  "Report  on  Safety  Valves," 
by  James  Brownlee,  in  the  "Transactions  of  Engineers  and  Ship- 
builders in  Scotland,"  Vol.  XVIII.  (Also  contained  in  London 
Engineering,  December  4  and  11,  1874.)  Table  I.  is  made  up 
from  the  data  of  two  sets  of  tests,  in  the  first  of  which  the  higher 
pressure  was  constant  and  the  lower  pressure  varied ;  while  in  the 
second  the  lower  pressure  was  constant  and  the  higher  pressure 
varied. 

These  tests  are  quoted  here  because  the  results  illustrate  Napier's 
rules  for  the  flow  of  steam.  In  making  the  tests  the  weight  of 
flow  was  measured  and  the  velocity  of  flow  was  calculated  by 
theoretical  formulas.  It  will  be  noted  from  column  4,  of  the  first 
group,  that  the  weight  of  steam  discharged  increases  until  the  lower 
pressure  drops  to  58  per  cent*  of  the  higher  pressure,  after  which 
the  weight  remains  constant.  This  is  as  it  should  be  from  Napier's 
experiments,  which  showed  that  the  flow  is  proportional  to  the 
higher  pressure,  when  the  lower  pressure  is  not  over  %0  °f  the 
higher. 


*The  value,  58  per  cent,  quoted  in  the  table  was  first  used  by  Weisbach,  and  ia 
frequently  given  instead  of  6-10,  as  the  limiting  point  at  which  Napier's  rules  hold.  Both 
figures  are  only  approximate,  however,  and  vary  considerably  with  the  conditions,  as 
will  be  shown.  See  reference  to  this  in  the  first  chapter,  in  connection  with  the  dis- 
cussion of  the  flow  of  steam. 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


219 


TABLE  I. 

BROWNLEE'S  EXPERIMENTS  ox  THE  FLOW  THROUGH 
CYLINDRICAL  ORIFICES. 


' 

Absolute 
pressure 
in  boiler, 
Ib.  sq.  in. 

Pressure 
against 
.vhich  nozzle 
discharged, 
Ib.  sq  in. 

Calculated 
velccity  at 
throat  of 
nozzle  in 
ft.  per  sec. 

Flow  of 
steam  in  Ibs. 
per  minute 
per  sq.  in. 
of  orifice. 

1 

~ 

;j 

4 

75 

74 

230 

16.68 

15 

72 

401 

28.35 

75 

70 

521 

35.93 

First  Group. 
Constant  initial 
pressure.    Variable 
final  pressure. 

75 
75 
75 
75 

65 
60 
50 
45 

i         43  46    / 

749 
933 
12:>2 
1401 

48.38 
56.12 
64.0 
65.24 

75 

_)               1O  .  **U      ' 

1  53  per  ct.  ( 

1446.5 

65.3 

75 

15 

1446.5 

C5.3 

75 

0 

1446.5 

65.3 

165 

14.7 

1481 

140.46 

135 

14.7 

1412 

115.61 

115 

14.7 

146(5 

98.76 

Second  Group. 
Variable  initial 

90 

70 

14.7 
14.7 

1454 
1444 

77.94 
61.07 

pressure.    Constant 

50 

14.7 

1429 

44.06 

final  pressure. 

40 

14.7 

1419 

35.18 

30 

14.7 

1408 

26.84 

25.37 

I         14.7     1 

|  53  per  ct.  (" 

1401 

22.81 

In  the  second  group  the  higher  pressure  drops  instead  of  being 
constant  as  before,  and  as  the  lower  pressure  is  in  each  case  less 
than  %0  of  the  higher  pressure,  the  flow  should  decrease  at  the 
same  rate  that  the  higher  pressure  drops,  which  is  actually  the  case. 

Another  lesson  to  be  learned  from  Brownlee's  table  is  that  the 
velocity  of  flow  is  practically  constant  when  the  lower  pressure  is 
%0  or  less  of  the  upper  pressure.  In  column  3,  the  velocity  is  cal- 
culated to  be  from  about  1,400  to  1,480  feet  per  second  when  the 
lower  pressure  does  not  exceed  %0  of  the  higher  pressure.  A 
value  of  1,450  feet  is  a  fair  average,  and  sometimes  it  is  said  in 
round  numbers  that  steam  will  discharge  from  either  a  converging 
or  straight  nozzle  at  the  rate  of  1,450  feet  per  second  when  the 
lower  pressure  is  %0  or  IGSS  of  the  upper. 

Flow  Through  Cylindrical  Nozzles. — Experiments  were  made 
by  L.  H.  Kunhardt,  class  of  1889,  Massachusetts  Institute  of  Tech- 
nology, under  the  direction  of  Prof.  C.  H.  Peabody  upon  the  flow 


220 


STEAM  TURBINES 


of  steam  through  tubes  or  mouthpieces  %  inch  in  diameter.  There 
were  three  mouthpieces  tested  having  inlets  rounded  with  a  radius 
of  1  inch  and  straight  sections  of  *4,  Y*  and  ll/2  inches,  respec- 
tively. The  experiments  were  conducted  to  find  the  weight  of 
steam  discharged  for  different  differences  of  pressure.  The  flow 
was  then  calculated  by  a  theoretical  formula  based  on  the  principles 
of  thermodynamics,  and  the  results  compared,  which  gave  the 
probable  coefficient  of  flow  for  the  three  mouthpieces.  The  flow 
was  also  calculated  by  Napier's  formula. 

When  under  test  the  mouthpieces  were  screwed  into  a  brass 
partition  between  two  cast-iron  reservoirs,  in  the  first  one  of  which 
the  steam  was  maintained  at  a  constant  initial  pressure,  and  in  the 
second  the  pressure  was  varied.  The  pressure  in  the  tube  was 
found  by  drilling  into  the  tube  at  the  middle  of  the  straight  section 
of  each  mouthpiece. 

TABLE   II. 
FLOW  THROUGH  SHORT  TUBES  WITH  ROUNDED  INLETS. 


05 

1 

Steam  pressures 
(gauge),  Ib.  sq.  in. 

Ratio  of 
pressures, 
absolute. 

Flow  in  pounds 
per  hour. 

a 

^ 

to 

aT 
,O 

t 

£ 

0) 

3  + 

f  i  . 

a 

^ 

a 

>>  ,  2 

&<* 

-2 

% 

-<-» 

£  *":  t^ 

.ti  ^  t? 

§ 

,Q 

"o  § 

"o 

(H 

m 

O 

% 
8 

c 

O 

of? 

ol^ 

| 

«1d 

-oS^ 

1? 

Numbe 

5 

bo 
a 

S 

'3 

Is 

G 

£ 

Pressui 

*  a  g 

C  a;  o> 

|§§ 

ft 

X 
0) 

>>' 

M 

Calcula 
theoret 
equatio 

"3  ^"5 
o'g.2 

Coeffici 
item  7 

£S» 

5ss 

1 

2 

3 

4 

5 

6 

7 

8 

9 

10 

1 

1.5 

74.1 

14.8 

41.2 

0.33 

0.63 

221.0 

217.0 

2-24 

1.018 

2 

1.5 

71.0 

13.2 

39.6 

0.33 

0.63 

213.0 

207.8 

215 

1.025 

3 

1.5 

72.6 

19.7 

40.6 

0.39 

0.63 

216.0 

211.4 

220 

1.022 

4 

1.5 

75.9 

20.4 

42.6 

0.39 

0.63 

228.0 

219.3 

227 

1.040 

5 

1.5 

71.9 

24.5 

40.6 

0.45 

0.64 

213.0 

209.7 

218 

1.016 

6 

0.5 

72.8 

14.8 

39.0 

0.34 

0.61 

225.0 

213.6 

221 

1.053 

7 

0.5 

72.1 

20.4 

38.8 

0.40 

0.62 

223.5 

211.7 

219 

1.056 

8 

0.5 

72.6 

24.7 

39.0 

0.45 

0.62 

223.0 

213.1 

220 

1.046 

9 

0.5 

73.1 

29.9 

39.2 

0.51 

0.62 

225.5 

213.0 

222 

1.054 

10 

0.25 

72.6 

24.8 

38.1 

0.45 

0.58 

225.0 

213.5 

220 

1.054 

11 

0.25 

72.6 

19.9 

36.1 

0.40 

0.53 

225.0 

213.5 

220 

1.054 

12 

0.25 

72.7 

14.9 

33'.2 

0.34 

0.58 

2:27.0 

213.0 

220 

1.066 

13 

0.25 

126.3 

27.8 

69.0 

0.29 

0.59 

3.58.8 

338.9 

355 

1.058 

14 

0.25 

125.0 

40.8 

67.9 

0.40 

0.60 

355.0 

334.8 

352 

1.060 

EXPERIMENTS  ON  THE  FLO  W  OF  STEAM  221 

It  is  to  be  observed  that  the  actual  flow  is  larger  than  that  cal- 
culated by  the  theoretical  equation.  This  equation  is  based  on  the 
assumption  that  no  heat  is  lost  to  or  given  up  by  the  tube.  But  it 
is  evident  that  some  heat  must  have  been  conducted  through  the 
walls  of  the  nozzle  from  the  hot  steam  in  the  upper  chamber  to  the 
steam  passing  through  the  tube,  and  this  may  explain  why  the  co- 
efficient of  flow  is  greater  than  unity  in  each  case.  The  results  in- 
dicate that  this  transmission  of  the  heat  was  more  than  enough  to 
overcome  any  loss  from  friction  in  the  tube. 

For  the  longest  tube  Napier's  rule  gives  results  greater  than  the 
actual  weight  of  steam  passing  through  the  nozzle,  and  for  the 
short  tubes  the  results  calculated  by  Napier's  rule  fall  below  the 
actual.  This  shows  the  greater  effect  of  friction  in  the  longest 
tube.  The  largest  discrepancy  between  the  actual  results  and  those 
given  by  Napier's  formula  is  about  3  per  cent. 

The  pressure  in  the  tube  ranges  from  .58  to  .64  of  the  upper  abso- 
lute pressure  and  is  more  for  the  long  tube  than  for  the  shorter 
ones,  and  is  also  slightly  more  for  the  high  pressure  tests  with 
the  short  nozzle  than  for  the  low  pressure  tests.  The  question  of 
the  pressure  of  steam  at  or  near  the  throat  of  the  nozzle  has  an  im- 
portant bearing  on  the  design  of  a  nozzle.  These  tests  and  others 
yet  to  be  quoted  show  that  the  throat  pressure  does  not  vary  widely, 
and  that  it  lies  between  .5  and  .7  of  the  upper  absolute  pressure,  or 
at  a  mean  of  about  .6.  Theoretically  it  should  be  .58  of  the  upper 
pressure. 

Experiments  on  the  Discharge  of  Steam  Through  Orifices,  by  Strick- 
land L.  Kneass.* 

These  experiments  were  completed  in  1890  at  the  works  of 
William  Sellers  &  Co.,  Inc.,  Philadelphia,  in  connection  with  their 
steam  injector  work.  Their  object  was  to  determine  the  behavior 
of  steam  within  a  discharging  nozzle,  and  the  extent  to  which  the 
terminal  velocity  is  affected  by  changes  in  the  proportion  of  the 
tube.  The  nozzles  tested  were  8mm.  (.31496  inch)  internal 
diameter  at  the  throat,  and  34mm.  (1.34  inches)  long.  The  other 
dimensions  of  the  nozzles,  however,  were  varied.  The  tubes  were 
connected  to  the  steam  supply  by  a  2-inch  pipe  and  care  was  taken 

•Proceedings  of  the  Engineers'  Club  of  Philadelphia,  July,  1891,  from  which  the 
following  abstract  was  prepared. 


222 


STEAM  TURBINES 


to  secure  dry  steam.  In  order  to  determine  the  pressures  within 
the  nozzle  seven  small  holes  were  drilled  equal  distances  apart  in 
the  walls  of  the  nozzle,  commencing  at  the  point  where  the  curve 
of  approach  becomes  tangent  to  the  cylindrical  barrel  of  the  tube. 
Each  of  these  apertures  had  gauge  connections  at  the  outer  end, 
and  the  holes  not  in  use  were  closed  by  plugs.  (See  Fig.  1.)  A 
small  searching  tube  was  used  for  finding  the  internal  pressure  of 
the  jet  at  points  beyond  the  end  of  the  nozzle.  The  tube  was  closed 
at  one  end  and  had  holes  drilled  in  one  side  near  the  end.  The 


SYPHON  AND  GAUGE 


Fig.  1.     Method  of  Measuring  Pressures. 

other  end  of  the  tube  was  connected  with  a  gauge,  and  by  placing 
the  tube  concentric  with  the  axis  of  the  nozzle  and  sliding  it  to 
different  positions  the  pressures  could  be  determined.  Since  the 
tube  changed  the  relation  of  the  areas  of  the  different  sections  of 
the  nozzle  in  which  it  was  inserted,  the  nozzle  was  made  of  pro- 
portionately larger  diameter  to  compensate  for  this,  when  the  tube 
was  used. 

Mr.  Kneass  was  the  first  in  this  country  to  make  systematic  in- 
vestigation of  steam  discharge  through  nozzles.  The  questions  of 
internal  pressures,  the  relation  between  length  and  terminal 
velocity,  and  the  taper  and  shape  of  nozzles  were  carefully  gone 
into. 

Explanation  of  Diagram. — The  diagram,  Fig.  2,  shows  a  longi- 
tudinal cross  section  of  each  of  the  five  nozzles  tested.  The  vertical 
lines,  I,  2,  2,  4,  etc.,  cutting  each  nozzle  pass  through  the  points 
at  which  the  pressures  were  measured.  Below  each  nozzle  section 
are  plotted  pressure  curves  passing  through  points  corresponding 


•QN003S  H3d  l~33d  Nl  A1IOCM3A 


•QN003S  U3d  J. 


224  STEAM  TURBINES 

to  the  observed  pressures,  while  above  the  nozzle  sections  are 
velocity  curves.  The  points  for  these  latter  were  calculated  by  first 
plotting  the  observed  pressures  and  volumes  of  a  unit  weight  of 
steam,  and  drawing  a  curve  through  the  points.  The  area  under 
the  curve  represents  the-  energy  in  foot  pounds  producing  the 
velocity  of  the  steam,  just  as  an  indicator  card  represents  work 
done.  The  velocity  corresponding  to  this  area  was  then  found  by 
the  formula 


taken  from  Rankine's  "The  Steam  Engine,"  page  298,  the  U  in 
the  formula  representing  the  work  done  by  unit  weight  of  steam  if 
admitted  to  a  cylinder  at  the  initial  pressure,  expanded  adiabatic- 
ally,  and  expelled  at  the  terminal  pressure.  Theoretically,  at  least, 
the  work  done  by  steam  in  giving  itself  velocity  in  a  non-conduct- 
ing nozzle  should  be  equal  to  the  work  done  by  steam  in  an  engine 
cylinder  under  like  pressure  ranges. 

Tests  upon  Nozzle  No.  I.  —  The  cylindrical  nozzle  was  first 
tested  and  the  pressure  curves  show  a  gradually  decreasing  pres- 
sure, steeper  at  high  pressures  than  at  low.  The  pressure  at  the 
first  hole  bears  nearly  a  constant  relation  to  the  initial  pressure, 
while  the  ratio  of  terminal  to  initial  pressure  falls  slightly  as  the 
initial  pressure  decreases.  These  points  are  shown  in  the  fol- 
lowing : 

Initial  Ratio  of  Ratio  of 

'  Steam  Throat  Pressure  Terminal  Pressure 

Pressure.  to  Initial  Pressure.  to  Initial  Pressure. 

P  P1  P* 

P  P 

120  0.700  0.585 

90  0.699  0.543 

60  0.666  0.533 

30  0.666  0.533 

20  0.657  0.529 

The  action  of  the  steam  and  cause  of  inefficiency  are  apparent. 
Taking,  for  example,  the  120-pound  line,  we  find  that,  while 
traversing  the  nozzle  the  pressure  is  reduced  only  from  80  to  64 
pounds  and  is  discharged  at  that  pressure  into  the  atmosphere, 
losing  nearly  66  per  cent  of  its  velocity-producing  power.  At  60 
pounds  initial  pressure  the  velocity  line  is  a  little  higher  than  at 
120  pounds. 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM  225 

Tests  ivith  Nozzle  No.  2. — In  this  nozzle,  diverging  with  a 
straight  taper  of  1  in  10  in  diameter,  the  throat  pressure  falls  to 
about  %0  of  the  initial  and  is  very  nearly  the  same  as  the  terminal 
pressure  in  the  cylindrical  tube.  With  high  initial  pressure  the  ex- 
pansion line  approaches  the  hyperbolic  curve  and  discharges  very 
close  to  the  atmospheric  line,  while  the  GO-pound  curve  falls  even 
into  the  vacuum  lines  and  issues  from  the  tube  in  a  jet  that  con- 
tracts to  a  smaller  diameter  than  that  of  the  orifice.  This  phenome- 
non occurs  with  the  30-pound  and  20-pound  lines  within  the  con- 
fines of  the  nozzle  and  the  pressure  at  the  mouth  is  greater  than 
that  at  the  fifth  orifice.  This  shows  that  for  these  lower  pressures 
there  is  a  loss  in  efficiency  in  going  beyond  this  point  of  maximum 
expansion,  and  that  the  discharging  jet  would  have  higher  terminal 
velocity  if  the  tube  were  made  shorter.  When  the  jet  expands 
within  the  nozzle  to  a  pressure  lower  than  15  pounds  absolute, 
there  will  be  a  contraction  of  the  jet  after  leaving  the  tube,  due 
to  the  pressure  of  the  atmosphere.  These  important  facts  were 
probably  first  noted  in  this  series  of  experiments. 

An  experiment  with  the  1  in  10  tube,  before  and  after  reaming 
the  discharge  end  with  steeper  tapers,  showed  that  such  change  in 
the  angle  of  divergence  produced  no  effect  upon  the  steam  dis- 
charge of  the  inlet  half  of  the  tube,  as  the  curves  of  velocity  and 
pressure  are  identical  for  the  inlet  end  of  the  tube  under  both 
conditions.  The  experiment  has  also  been  tried  of  gradually 
shortening  the  tube  by  cutting  sections  off  the  outlet  end,  and  it 
was  found  that  the  pressures  nearer  the  inlet  were  not  affected  by 
the  alteration. 

Design  of  Nozzle  No.  5. — In  all  the  straight  nozzles  there  is  loss 
of  energy,  due  to  fluid  friction,  owing  to  changes  in  velocity. 
Mr.  Kneass  reasoned  that  these  changes  in  velocity  should  cause 
less  loss  and  internal  friction  if  the  acceleration  of  the  jet  were 
made  uniform.  Nozzle  No.  5  was  designed  to  give  constant  ac- 
celeration. The  pressure  at  the  throat  was  assumed  at  its  lowest 
theoretical  value  and  the  terminal  velocity  at  the  highest  attainable 
in  expanding  from  120  pounds  to  the  atmosphere.  The  tube  was 
divided  into  seven  parts  and  the  acceleration  calculated.  Then,  in 
order  to  divide  the  work  evenly,  the  expansion  was  made  uniform 
and  from  these  data  the  area  of  the  tube  at  each  given  section  was 


226  STEAM  TURBINES 

determined.  As  shown  by  the  diagram,  this  tube  gives  superior 
results. 

Deductions  from  the  Velocity  Lines. — It  will  be  noticed  that, 
notwithstanding  the  marked  differences  in  initial  pressure,  ranging 
from  30  to  120  pounds,  gauge,  the  velocity  curves  for  the  several 
nozzles,  and  particularly  for  nozzle  5,  fall  very  close  together.  In 
the  latter  nozzle  the  velocities  at  any  given  point  are  practically  the 
same  for  all  the  pressures  used.  Mr.  Kneass  finds  by  analyzing 
formulas  of  Rankine  for  the  flow  of  steam  that  if  the  ratio  of  the 
throat  to  the  initial  pressure  were  the  same  under  all  conditions, 
there  would  be  a  lower  velocity  for  30  than  for  120  pounds  initial 
pressure  at  any  given  point  in  the  nozzle.  It  will  be  seen  from 
Table  III.,  however,  where  this  ratio  is  given  for  the  different  noz- 
zles and  for  the  different  initial  pressures,  that  its  value  decreases 
with  the  initial  pressure.  In  other  words,  the  throat  pressure  is  a 
smaller  percentage  of  the  initial  pressure  when  steam  is  being  used 
at  a  low  initial  pressure  than  when  used  at  a  high  initial  pressure. 
This  condition,  apparently,  affects  the  velocity  at  any  point  in  the 
nozzle  in  a  way  partially  to  neutralize  the  effect  of  the  initial  pres- 
sure upon  the  velocity.  This  is  shown  both  by  the  formulas  and 
the  tests.  For  example,  in  nozzle  5,  taking  the  actual  throat  pres- 
sures as  found  by  the  tests,  the  velocities  at  point  4  for  the  different 
initial  pressures  are  found  by  calculation  to  vary  only  by  4  per  cent. 

In  Table  IV.  are  the  velocities  of  the  steam  for  the  different 
points  in  the  several  tubes  as  calculated  from  the  experimental  re- 
sults. The  variation  of  the  velocity  from  a  constant  value  for  all 
the  pressures  in  the  table  is  small.  Mr.  Kneass  states  that  for 
practical  purposes  the  velocity  may  be  considered  constant  and  a 
simple  formula  derived  for  the  weight  of  steam  discharged  through 
an  orifice  in  a  given  time  which  is  more  convenient  to  apply  than 
the  thermodynamic  equation.  It  is  based  on  the  assumption  of 
constant  velocity  of  discharge  and  is  to  be  applied  where  the  ratio 
of  final  to  initial  absolute  pressures  does  not  exceed  0.6. 

Let  W=weight  in  pounds  discharged  per  second, 
a=area  of  orifice  in  square  inches, 

d=weight  of  1  cubic  foot  of  steam  in  its  initial  condition. 
Then,  PF=6.19  ad.  (3) 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 
TABLE  III. 

RATIO  OF  THROAT  PRESSURE  TO  INITIAL  PRESSURE. 
(Both  in  absolute  pressures.) 


227 


Shape  of  nozzle. 

Initial 

pressure, 
gauge 
Ibs.  sqr.  in. 

Cylindrical. 

1  in  10. 

Iin6. 

1  in  6  and 
1  in  10— 

Iin5. 

Special 
curve. 

mean. 

120 

.703 

.610 

.606 

.608 

.539 

.517 

90 

.689 

.589 

.582 

.589 

.498 

.520 

60 

.665 

.558 

.558 

.560 

.501 

.508 

30 

.664 

.552 

.541 

.552 

, 

.507 

20 

.654 

.524 

.553 

.520 

.430 

.... 

10 

.676 

.510 

.545 

.506? 

.496 

.... 

TABLE  IV. 
STEAM  VELOCITIES  IN  NOZZLES. 


Shape  of  tube. 

Initial 
pressure 
(gauge). 

Velocity  in  feet  per  second. 

Sections  of  nozzle. 

1 

2 

3 

4 

5 

6 

7 

Cylindrical. 

120 
60 

1195 
1261 

1C52 
1291 

1349 
1416 

1368 
1445 

1400 
1513 

1464 
1560 

1500 
1589 

Straight  taper,  1  in  10. 

120 
90 
60 
30 

1386 
1541 
1532 
1517 

1?99 
1<H6 

1882 
1926 

2081 
2246 
2168 
2047 

2250 
23?8 
2272 
2235 

2:389 
2166 
2384 
2341 

2495 
2601 
2459 
2235 

2668 
2707 
2601 
2143 

Straight  taper,  1  in  6. 

12(1 
90 
60 
30 

1400 
1483 
1310? 
1513 

1927 
1822 
1960 
1927 

2328 
2312 
2389 
2119 

2500 
2457 
2476 
2380 

2669 
2664 
2484 
2479 

2765 
2707 
2712 
2158 

2855 
2867 
2652 
2139 

Straight  taper,  1  in  5. 

120 
80 

eo 

30 

1504 
1710? 
1638 
1599 

2091 
2062 
2071 
2042 

2466 
2420 
2447 
2408 

2632 
2526 
2483 
2216 

2801 
2646 
2678 
2235 

2910 
2768 
2755 
2206 

2975 
2827 
2505 
2158 

Special  curve,  uniform 
acceleration.  Nozzle 
5. 

120 
90 
60- 
30 

1662 
1681 
1696 
1626 

2065 
2069 
2023 
1956 

2272 
2267 
2216 
2187 

2476 
2462 
2457 
2466 

2657 
2563 
2592 
2331 

2746 
2663 
2717 
2206 

2890 
2823 

2H81 
2148 

Mean  tor  nozzle  5  (con- 
tinuous expansion). 

1666 

2028 

2235 

2465 

2604 

2708 

28134 

228  STEAM  TURBINES 

Example. — a=l ;  initial  pressure=100  pounds  absolute; 
d=.22Vl  (from  steam  table).  Then,  W=6.19X1X.2271=1.405 
pounds  per  second.  This  is  to  be  compared  with  the  same  example 
solved  by  equation  (1),  this  chapter,  according  to  Napier's  rule. 

Experiments  on  Steam  Jets  by  Walter  Rosenhain.* 

Description  of  Apparatus. — In  this  series  of  experiments  both 
cylindrical  and  diverging  nozzles  were  used,  and  the  velocity  of  dis- 
charge was  measured  by  the  ingenious  plan  of  weighing  the  re- 
action of  the  jet  upon  the  nozzle  by  means  of  a  special  apparatus. 
Knowing  the  reaction  of  the  jet  in  pounds  and  the  weight  of  steam 
discharged  in  a  given  time,  the  velocity  could  be  accurately  deter- 
mined.f 

In  Fig.  3  is  a  diagram  of  the  apparatus.  A  vertical  tube,  F,  is 
joined  to  a  steam  pipe,  D,  and  supports  at  its  lower  end  a  cylindrical 
chamber,  H,  which  has  a  circular  opening  for  the  nozzle.  The 
tube,  F,  is  of  bicycle  tubing,  which  has  sufficient  flexibility  to  allow 
the  cylindrical  chamber  to  swing  freely  about  the  point  of  sup- 
port at  D.  The  reaction  of  the  jet  is  measured  by  weights  placed 
in  the  scale  pan,  S,  carried  by  a  cord  passing  over  the  pulley,  P, 
and  attached  to  the  horizontal  arm  at  K.  A  pointer  indicates  the 
movement  of  the  arm  and  chamber,  H,  on  the  scale  shown  above 
the  arm.  Pulley  P  is  a  finely  finished  steel  disk,  with  ball  bearings. 
A  gauge  registers  the  pressure  of  the  steam  in  H.  In  the  tests  the 
pressure  of  the  steam  supplied  the  apparatus  was  controlled,  either 
by  varying  the  boiler  pressure  or  by  throttling  the  steam  after  it 
left  the  boiler.  After  the  observations  of  reaction  were  made,  the 
discharge  was  measured  in  pounds  per  second  under  as  nearly  the 
same  conditions  as  possible,  the  steam  being  conducted  to  a  surface 
condenser  and  weighed. 

The  formula  for  the  velocity  of  discharge  is  based  on  the  old 

•Proceedings  of  The  Institution  of  Civil  Engineers,  London,  Vol.  CXL.,  1900. 

fA  series  of  articles  in  London  Engineering,  1872,  upon  "Experiments  and  Re- 
searches of  the  Efflux  of  Elastic  Fluids,"  by  Wilson,  describes  a  similar  method  for 
measuring  the  velocity  or  a.  jet.  Wilson's  apparatus  was  elaborate  and  his  tests  were 
exhaustive,  but  the  results  are  not  of  value  for  the  present  purpose.  Another  method 
has  been  used  by  Strickland  L.  Kneass,  who  arranged  the  nozzle  to  discharge  against 
a  delicately  balanced  parabolic  target.  The  target  turned  the  jet  through  an  angle  of 
90  degrees  and  the  pressure  against  the  target  should  give  a  result  equivalent  to  the 
reaction  of  a  swinging  nozzle. 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


229 


principle  that  action  and  reaction  are  equal, — the  accelerating  force 
of  the  jet  is  equal  to  the  reaction  of  the  jet  upon  the  nozzle  and  its 
chamber,  H, — and  may  be  derived  thus :  Suppose  a  force,  F,  act- 
ing as  a  constant  pull  or  push  on  a  free  body,  to  give  the  body  a 
velocity  of  V  feet  per  second  at  the  end  of  one  second.  Then,  since 


Fig.   3.     Method  of  Measuring  Velocity  of 
Flow. 

we  know  that  gravity,  acting  on  the  same  body  with  a  constant 
force  of  W  pounds,  equal  to  the  weight  of  the  body,  would  produce 
a  velocity  of  32.2  feet  per  second  at  the  end  of  one  second,  we  have, 


P:W=V:g,  or, 

Fg 
V=—. 

W 


(4) 


230 


STEAM  TURBINES 


In  applying  this  formula  to  the  steam  jet,  we  have 

['^velocity  in  feet  per  second, 

F=  reaction  in  pounds, 

<g-=acceleration  of  gravity  (=32.2), 
£F— pounds  steam  discharged  per  second. 

TABLE  V. 
EXPERIMENTAL  NOZZLES. 


No. 

Least 
Diameter. 

Greatest 
Diameter. 

Length. 

Taper. 

/. 
II. 

//.    A 

0.1873 
0.1840 
0.1866 

6!  287* 
0.1866 

Z.\ 
0.5 

Thin  plate 
1  in*0 

II.    B 
III. 
III.    A 
III.    B 
IV. 

0.1849 
0.188-2 
0.1882 
0.1882 
0.1830 

0.287 
0.368 
0.255 
0.241 
0.255 

1.6 
2.16 
0.79 
0.64 
2.16 

linSO 
1  in  13 
Iinl2 
1  in  12 
1  in  30 

The  Nozzles  Used  in  the  Tests  are  represented  in  Fig.  4,  and 
their  dimensions  are  tabulated  in  Table  V.  The  nozzles  were  all 
made  with  a  throat  diameter  as  nearly  as  possible  3-16  inch  in 
diameter,  the  exact  dimensions  being  given  in  the  table.  This  is  the 
diameter  of  the  nozzles  of  a  De  Laval  5-horse  power  turbine.  The 
dimensions  of  De  Laval  nozzles  of  this  size  for  several  different 
pressures  are  given  in  Table  VI.,  which  shows  the  tapers  to  vary 
from  about  1  in  17  to  1  in  27.  Guided  by  this,  the  tapers  of  the  ex- 
perimental nozzles  were  made  1  in  12,  1  in  20,  and  1  in  30,  which 
gave  wide  enough  latitude  for  the  tests  and  yet  kept  reasonably 
close  to  proportions  that  had  proved  satisfactory  in  practice. 

TABLE  VI. 
DE  LAVAL  NOZZLES  FOR  5-H.  P.  TURBINE. 


Pressure  - 
Ibs.  sq.  in. 

Least 
Diameter. 

Length. 

Taper. 

Inl. 

Inl. 

136 

0.157 

1.57 

1  in  17.  4 

105 

0.163 

1.57 

1  in  21.  4 

100 

0.197 

1.57 

1  in  19.0 

60 

0.230 

1.57 

1  in  29.0 

58 

0.256 

1.57 

1  in  26.6 

EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


231 


Preliminary  Experiments  in  Weight  and  Velocity  of  Flow. — 
As  evident  in  Fig.  4,  nozzles  II.  A  and  II.  B  are  simply  the  two 
separate  parts  of  nozzle  //.,  which  has  a  well  rounded  inlet  and  a 
taper  in  the  diverging  part  of  1  in  20.  Preliminary  experiments 
were  run  with  nozzles  /.,  II. ,  II.  A,  and  II.  B,  and  the  results  are 
plotted  in  the  diagrams  in  Figs.  5  and  6.  Referring  to  the  former, 


w,. 


THIN   PLATE 


II   CONVERGING  AND  DIVERGING 


II  A 

CONVERGING  OR 
CYLINDRICAL 


-III  B— 
—III  A- 


III 

DIVERGING,  SLIGHTLY 
ROUNDED  INLET 


TAPER  1  IN  30 


IV 

DIVERGING,  SLIGHTLY 
ROUNDED  INLET 


Fig.   4.      Nozzles   Used   by   Rosenhain. 


which  gives  the  weight  discharged  per  second  for  different  pres- 
sures, nozzle  //.,  having  an  easy  inlet  and  an  expanding  outlet, 
gives  the  greatest  discharge.  Next  comes  the  converging  mouth- 
piece, //.  A,  which  indicates  that  the  inlet  section  of  nozzle  II.  is 
more  important  than  the  outlet  section  in  its  effect  upon  the  quan- 
tity of  discharge.  The  position  of  //.  B  so  far  below  7.  seems  to 
show  that  the  sharp  inlet  is  unsuited  to  passing  a  large  quantity  of 
steam  through  an  expanding  nozzle. 

The  velocity  curves  in  Fig.  6,  on  the  other  hand,  show  that  while 


232 


STEAM  TURBINES 


the  quantity  of  steam  passed  by  a  nozzle  depends  very  considerably 
on  the  shape  of  the  inlet,  the  velocity  on  leaving  the  nozzle  depends 
more  on  the  shape  of  the  outlet  portion.  This  points  to  the  con- 
clusion that  the  density  of  the  steam  at  the  throat  of  the  nozzle  de- 
pends upon  the  shape  of  the  inlet  and  that  this  density  is  greater 
with  a  well-rounded  inlet  than  with  a  nozzle  having  a  sharp  inner 
edge. 

This  accounts  for  the  most  conspicuous  feature  of  this  set  of 
velocity  curves,  viz.,  that  up  to  a  pressure  of  about  80  pounds  per 


.07 

I.« 

B 

£ 


DISCH 

S 


INITIAL_PRE8SURE,  LB9.  SQ.  IN.,  GAUGE. 
50         70          90         110        130        15 


17Q 190_210 


II. 

1IA 


Fig.  5. 

square  inch  the  greatest  velocity  is  attained  by  a  jet  from  an  orifice 
in  a  thin  plate  and  that  above  100  pounds  per  square  inch,  //.  Bt 
having  a  sharp  inlet,  gives  a  greater  velocity  than  //.,  which  has 
a  rounded  inlet  and  the  same  outlet.  Apparently  the  rounded  inlet 
admits  a  greater  weight  of  steam  to  the  narrowest  section  than  the 
nozzle  can  deal  with  efficiently. 

The  reason,  of  course,  why  the  diverging  nozzles  give  greater 
velocities  at  the  higher  pressures  than  the  thin  plate  or  the  con- 
verging nozzle  is  that  expansion  is  not  complete  in  the  two  latter 
at  the  higher  pressures  and  there  is  wasted  energy. 

Comparing  the  dimensions  of  nozzle  II.  B  with  the  dimensions 
of  the  De  Laval  nozzles,  it  is  found  to  lie  midway  between  the  ox- 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


233 


tremes,  and  hence  two  other  nozzles  were  designed,  ///.  and  IV., 
having  tapers,  respectively,  of  1  in  12  and  1  in  30,  as  previously 
stated.  These,  however,  were  made  with  inlet  rounded  slightly, 
with  small  radius,  instead  of  with  a  large  radius,  as  in  nozzle  II.; 
this  in  view  of  the  fact  that  the  preliminary  tests  had  shown  that 
when  the  inlet  was  considerably  rounded  the  velocity  of  flow 
would  not  be  as  great. 

Tests  with  Nozzle  III. — In  order  to  secure  information  upon 
the  best  ratio  of  outlet  to  throat  diameters  of  a  steam  nozzle,  tests 


3000 


J» 


INITIAL  PRESSURE,  LBS.  SO.  IN.,  GAUGE. 

40    60    80    100    120   140    160 


II//IIB 


Fig.  6. 


were  run  with  nozzle  ///.  at  its  original  length  of  2.16  inches  and 
then  it  was  cut  down,  first  to  .79  inches  and  finally  to  .64  inches 
long.  Nozzle  ///.  B  discharged  the  greatest  quantity  of  steam, 
nozzle  ///.  slightly  less  and  ///.  A  the  least.  The  discharge  curves 
keep  very  close  together,  however,  and  it  is  not  thought  necessary 
to  reproduce  them  herewith,  as  their  mean  values  very  nearly  co- 
incide with  curve  //.  A,  Fig.  5,  for  the  converging  nozzle.  The 
velocity  curves  are  given  in  Fig.  7,  which  shows  nozzle  ///.  A  to 
be  the  most  efficient.  Comparing  the  curves  in  Fig  7,  it  seems 
probable  that  in  nozzle  III.  B  expansion  was  insufficient  and  that 


234 


STEAM  TURBINES 


in  nozzle  III.  it  was  too  great.  It  may  be  supposed  that  in  the  lat- 
ter the  steam  expanded  down  to  atmospheric  pressure  and  then 
flowed  through  the  remaining  portion  of  the  nozzle  much  as  an 
incompressible  fluid  would  do,  the  steam  increasing  in  section  at 
the  expense  of  the  velocity.  (See  velocity  diagrams,  also,  in 
Fig.  2.) 

Nozzle  ///.  A,  with  1  in  12  taper,  is  roughly  comparable  with 
nozzle  //.  B,  which  showed  up  the  best  of  the  1  in  20  tapers,  and 
from  their  velocity  curves  there  appears  to  be  a  small  difference 
in  favor  of  the  1  in  12  taper. 


win0        2°       * 

INITIAL  PRESSURE,  LB8.  8Q.  IN.  ,  Gf 

)         60         80        100        120       U 

UGE-                             „. 

o     160     IM   J2PIIIA 

^-"            _^-III3 

VELOCITY  IN  FT.  PER  SEC. 

/ 

/ 

^^ 

/ 

/, 

/^  • 

/ 

/ 

/ 

y 

_      / 

'      / 

/ 

/ 

/ 

/    / 

/ 

/ 

/ 

/ 

// 

/ 

// 

/ 

// 

/ 

/ 

f 

ll 

1 

II 

/ 

1200 

11  ii  in 

/ 

1 

V 

1 

Fig.    7. 

Tests  with  Nozzle  IV. — It  will  be  noted  from  Table  V.  that 
nozzle  IV.  is  so  designed  as  to  be  directly  comparable  with  nozzles 
///.  and  ///.  A.  It  has  the  same  length  as  nozzle  III.t  but  a 
smaller  outlet  diameter,  since  its  taper  is  1  in  30  instead  of  1  in  12. 
Its  outlet  diameter  is  the  same,  however,  as  of  nozzle  ///.  A,  but 
its  length,  of  course,  is  greater.  The  velocity  curves  for  these  three 
nozzles  are  plotted  in  Fig.  8.  Nozzle  ///.  gives  the  poorest  results, 
as  before,  in  diagram  in  Fig.  7,  and  ///.  A  gives  the  best  results,  as 
before,  while  nozzle  IV.  falls  between.  The  conclusion  has  already 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


235 


been  drawn  that  nozzle  ///.  carried  the  expansion  too  far,  which 
accounts  for  its  position  on  the  diagram ;  while  as  between  ///.  A 
and  IV.,  both  of  which  expand  the  steam  to  the  same  extent,  it  is 
very  evident  that  ///.  A,  which  has  the  taper  of  1  in  12,  is  much  the 
superior. 

After  experimenting  with  nozzle  IV.,  Rosenhain  measured  the 
flow  after  the  nozzle  had  been  shortened  by  y%  inch  at  a  time,  until 
it  was  only  .66  inch  long,  but  the  results  have  no  special  signifi- 
cance beyond  what  has  been  shown  by  the  other  tests. 


Ill  A 

3000 


77 


A 


2600 
2100 


-IV 
'III 


1000 


20    10    60    80    100   120   110   160   180   200 

INITIAL  PRESSURE.  LBS.  SQ.  IN.,  GAUGE. 

Fig.   8. 

Conclusions  from  Rosenhain's  Tests. — Rosenhain  concludes 
from  his  experiments  that  the  most  efficient  form  of  nozzle  de- 
pends upon  how  great  the  pressure  is.  Up  to  80  pounds,  gauge 
pressure,  an  orifice  in  a  thin  plate  is  the  most  efficient  form  used 
in  these  experiments,  but  this  does  not  imply  that  it  is  the  most 
efficient  form  which  can  be  used. 

For  high  boiler  pressures,  an  expanding  nozzle  with  inner  edge 
only  slightly  rounded  should  be  used. 

The  taper  should  not  be  very  different  from  1  in  12,  and  the 


236 


STEAM  TURBINES 


proper  ratio  of  greatest  and  least  diameters  is  given,  according  to 
present  results,  in  the  following  table : 

160 
Steam  pressure  in  pounds  per  square  inch 80 


Ratio  of  diameters. .  .  .1.26 


100 

1.26 

to 

1.33 


140 


to 
200 


1.36        1.36 


Rateau's  Experiments.* 

Professor  A.  Rateau,  of  Paris,  the  inventor  of  the  Rateau  steam 
turbine,  has  made  experiments  upon  the  escape  of  steam  through 
circular  orifices.  He  experimented  with  the  nozzles  and  the  orifice 
shown  in  Fig.  9,  the  diameters  of  which  are  as  follows : 

Nozzle.  Millimeters.  Inches. 

A  10.49  .412 

B  15.19  .598 

C  24.20  .954 

D  20.12  .793 


Fig.  9.     Nozzles  Used  by  Rateau. 

Tests  with  Final  Pressure  Less  than  .58  Initial  Pressure. — In 
this  series  a  large  number  of  tests  were  made,  and  a  few  data  and 
results  for  each  of  the  nozzles  have  been  selected  from  the  tables 
published  by  Rateau  and  grouped  in  Table  VII.  herewith.  The 
original  metric  units  and  their  English  equivalents  are  given.  In 
the  case  of  the  three  converging  nozzles  the  weight  of  steam  dis- 
charged depends  only  on  the  initial  pressure  and  is  not  affected  by 

*Paper  by  A.  Rateau  upon  "L'  Ecoulement  de  la  Vapeur  D'  Eau  par  Des  Tuyers 
et  des  Orifices"  in  the  "Annales  des  Mines/'  Paris,  January,  1902.  Since  this  abstract 
was  made  this  paper  has  been  translated  and  is  now  published  by  the  D.  Van  Nostrand 
Company,  New  York. 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


237 


TABLE  VII. 
FLOW  THROUGH  CONVERGING  NOZZLES  AND  ORIFICE  IN  FLAT  PLATE. 


Absolute  pressures. 

Flow  per 

Ratio  of  I  to  P. 

Initial  P. 

Final  p. 

s.cond  =  I 

vr  .-.__IQ 

Rcitio 

IX  OZZ1G. 

p 

N 

„• 

M 
fi 

„• 

.c) 

p 

0 

2; 

| 

O 

_g 

o 

O 

a 

CO 

0 
C 

j 

"co 

CO 

> 

i<5 

CO 

a 

CO 

'£ 

"3 

bo 

n 

bo 

S 

j^ 

3i 

c 

M 

J 

M 

O 

J 

* 

w 

r 

9  16 

130.30 

.231 

3.28 

081 

133.34 

1.89 

14.52 

.0145 

10.41 

148. 

.151 

2.145 

.'0/5 

149.SI6 

2.13 

14.40 

.0144 

^    J 

8.46 

120.40 

.131 

1.8-5 

.016 

123  32 

1.75 

14.57 

.0146 

A      -j 

Ii.l4 

153.40 

.151 

2.145 

.014 

159.1-2 

2.26 

14.31 

.0143 

10.74 

152.70 

4.89 

69.60 

.455 

133.96 

2.17 

14.^4 

.0112 

I 

10.43 

148.30 

5.99 

b5.30 

.571 

149.43 

».13 

14.3-2 

.0113 

r 

9.26 

131.85 

.144 

2.05 

.016 

132.43 

1.83 

1  !.3l 

.0143 

8.43 

119.85 

.144 

2.05 

.017 

122.03 

1.74 

1  L49 

.0145 

/?   J 

6.63 

95. 

.144 

2.05 

.022 

97.97 

1.89 

14.68 

.0147 

*»   1 

4.48 

63.70 

.039 

1.27 

.020 

67.01 

.95 

14.90 

.0149 

5.27 

74.9 

.339 

5.25 

.070 

76.52 

1.08 

14.52 

.0145 

5.39 

76.7 

.249 

3.54 

.046 

77.83 

1.11 

11.44 

.0144 

r 

2.44 

34.65 

.154 

2.'9 

.063 

36.39 

.517 

14.92 

.0149 

1.15 

13.33 

.127 

1.81 

.110 

17.43 

.248 

15.19 

.015-2 

1 

2.  '.13 

41.6) 

.171 

2.43 

.058 

43.31 

.616 

14.78 

.0.4* 

1 

4.03 

57.25 

.235 

3.3! 

.055 

53.30 

.829 

14.46 

.0145 

i 

2.97 

42.2) 

.934 

13.70 

.322 

43.76 

.6-21 

14.73 

.0147 

l 

1.19 

16.91 

.644 

9.16 

.541 

17.93 

.255 

15.11 

.0.51 

4.0! 

57.50 

.18) 

25.83 

.450 

51.92 

.739 

12.85 

.01-28 

3.86 

54.90 

.9)3 

14.12 

.257 

49.45 

.703 

1-2.82 

.0123 

2.57 

36.56 

.953 

13.62 

.371 

32.85 

.467 

12.78 

.0128 

z?   - 

2.36 

33.53 

1.38 

19.58 

.531 

27.13 

.385 

11.50 

.0115 

4.14 

53.80 

1.35 

19.23 

.321 

51.27 

.728 

12.40 

.0124 

3.81 

54.50 

1.83 

26.80 

.49) 

45.55 

.647 

11.88 

.0119 

2.93 

42.50 

1.59 

22.65 

.534 

35.18 

.530 

11.73 

.0117 

the  final  pressure.  This  is  shown  by  the  last  two  columns  of  the 
table,  which  give  the  discharge  per  kilogram,  or  pound,  initial 
pressure,  according  as  metric  or  English  units  are  taken.  These 
columns  express  the  ratio, 

/      Wt.  discharged  per  second  per  unit  area 
P  Initial  pressure  per  unit  area 

It  will  be  noted  that  this  ratio  is  very  nearly  a  constant  quantity 
except  for  the  orifice,  D,  in  the  thin  plate,  showing  that  flow 
through  the  latter  does  not  follow  the  same  law. 


z38  STEAM  TURBINES 

Rateau  plotted  the  results  of  these  tests,  using  as  ordinates  the 
ratios  I/P  and  as  abscissas  the  discharge  per  second  per  unit  area. 
He  then  plotted  the  theoretical  discharge  line,  using  a  formula  de- 
rived by  the  principles  of  thermodynamics,  in  order  to  compare  re- 
sults. The  differences  da  not  usually  exceed  two  per  cent.  The 
actual  discharge  was  slightly  in  excess  of  the  theoretical,  the  mean 
deviation  for  nozzle  A  being  .012  of  the  actual  discharge ;  for  noz- 
zle B,  .007,  and  for  nozzle  C,  .003  of  the  actual.  The  difference  in 
the  sizes  of  the  nozzles  apparently  had  no  marked  effect  on  the  re- 
sults. 

Formula  for  Weight  Discharged. — From  the  foregoing  it  is  evi- 
dent that  a  single  formula  may  be  employed  to  calculate  approxi- 
mately the  weight  discharged  for  all  pressures  within  the  limits  of 
the  tests,  provided  the  final  pressure  is  not  more  than  .58  of  the 
initial.  The  following  is  proposed,  the  form  of  which  is  derived 
by  theory,  but  the  constants  of  which  are  taken  from  the  tests : 

I=P  (15.26— .96  logP), 

in  which  /  is  the  flow  in  grammes  per  square  centimeter  per  second, 
and  P  is  the  initial  pressure  in  kilograms  per  square  centimeter. 
This  may  be  called  the  formula  of  maximum  discharge,  since  it 
gives  the  greatest  quantity  that  will  flow  through  the  nozzle  for  a 
given  pressure. 

In  English  units  it  reduces  to 

7=.001P    [15.26— .96    (log  P+log   .0703)], 

where  /  is  the  flow  in  pounds  per  square  inch  per  second  and  P  the 
initial  pressure  in  pounds  per  square  inch. 

Example. — Given,  P=6.6S,  metric ;  95,  English. 
In  metric  units, 

7=6.68  (15.26— .96X-8248) 
=6.68X14.47 
=96.6  grams  per  square  centimeter  per  second. 

In  English  units, 

7=.001X95  (15.26— .96X.825) 
=.095X14.47 
=1.374  pounds  per  square  inch  per  second. 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


239 


Tests  with  Final  Pressure  More  than  .58  of  Initial  Pres- 
sure.— Under  this  condition  the  discharge  depends  upon  the  final 
pressure  as  well  as  the  initial  pressure  and  the  results  may  be  rep- 
resented by  taking  the  ratio  of  the  measured  discharge  to  the 
maximum  calculated  discharge  /,  obtained  by  the  formulas  above. 


CURVES  FOR  DISCHARGE  OF  STEAM 

MAINLY  WHEN  FINAL  PRESSURE  IS  GREATER 
THAN  58$  OF  INITIAL  PRESSURE 


0.8  0.7  0.6 

RATIO  Or  FINAL  TO  INITIAL  PRESSURE 


Fig.  10. 

The  diagram,  Fig.  10,  shows  two  curves,  one  for  the  converging 
nozzles  and  one  for  the  thin  plate,  plotted  on  this  basis.  The  or- 
dinates  represent  the  ratio : 

Measured  discharge  when  p  is  more  than  .58  P 
Calculated  discharge  when  p  is  less  than  .58  P 

Where  p  is  the  final  and  P  the  initial  absolute  pressure.  The 
abscissas  give  the  ratio  p/P. 

Flow  Through  Orifice  in  Thin  Plate. — Different  laws  govern 
this  case.  The  discharge  does  not  become  the  maximum  for  final 
pressure  equal  to  or  little  less  than  .58  P,  but  it  increases  con- 


240 


STEAM  TURBINES 


stantly  as  p  lowers.  In  the  diagram,  Fig.  10,  are  the  relative  results 
up  to 

p/P=QA. 

By  taking  the  ratio  between  the  ordinates  of  this  curve  and  those 
of  the  experimental  curve  for  the  converging  mouthpieces,  the 
points  fall  on  a  straight  line  CD,  tangent  to  the  thin  plate  curve  at 
point  D,  which  is  on  the  ordinate  of  the  point  where  the  curve  AB 
joins  the  horizontal. 

Example. — Initial  pressure  absolute=10  kg.  sq.  cm.     Ratio  of 
final  to  initial  pressure=.  8.    From  the  diagram  we  find,  for  a  con- 


NOZZLE  No.  1 


NOZZLE  No.  2 
O^IS^DIAM. 


NOZZLE  No.  4 


No.  1   WITH   DIVERGENT  No.  2  WITH   DIVERGENT 

MOUTHPIECE  ADDED  MOUTHPIECE  ADDED 

TAPER  1   IN  23  TAPER  1   IN  23 

Fig.  11.  Nozzles  Used  by  Gutermuth. 

verging  nozzle,  that  the  ordinate  .8  cuts  the  curve  at  a,  and  this 
point  projected  to  the  left  gives  .82  for  the  ratio  of  discharge. 
Now  calculate  /  for  P=10,  and  multiply  the  result  by  .82,  which 
will  give  the  flow  through  the  converging  nozzle.  The  discharge 
through  the  thin  plate  can  be  calculated  in  the  same  manner,  by 
means  of  the  other  curve. 

Other  Experiments. 

Gutermuth's  Experiments  on  the  Outflow  of  Steam. — An  ac- 
count of  experiments  by  Professor  M.  F.  Gutermuth  of  the  Tech- 
nical High  School,  Darmstadt,  is  given  in  the  Journal  of  the 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


241 


American  Society  of  Naval  Engineers  for  May,  1904.  Nozzles, 
orifices  and  passages  of  various  forms  were  tested,  the  shapes 
shown  in  Fig.  11  being  those  best  adapted  to  turbine  work.  The 
tests  were  conducted  in  groups,  in  each  of  which  the  initial  pres- 
sure was  kept  constant  and  the  final  pressure  varied  to  give  the 
desired  pressure  differences.  This,  it  should  be  said,  is  the 
correct  way  in  which  to  obtain  comparative  results  and  it  is 
unfortunate  that  more  experimenters  have  not  adopted  the  same 
plan.  The  weight  and  velocity  of  steam  flowing  from  a  nozzle 
depend  more  upon  the  initial  pressure  than  upon  the  final  pressure, 

TABLE  VIII. 
COMPARISON  OF  FLOW  THROUGH  NOZZLES  SHOWN  IN  FIG.  11. 


Absolute 

Absolute 

Flow  in  kilograms  per  hour  —  actual. 

Ratio  of  flow. 

initial 

final 

pressure 
in  atmos- 
pheres. 

pressure 
in  atmos- 
pheres. 

Nozzle 
No.  I. 

Nozzle 
No.  II. 

Nozzle 
No.  III. 

Nozzle 
No.  IV. 

Nozzle 
I.  to 
nozzle  II. 

Nozzle 
III.  to 
nozzlelV. 

1 

2 

3 

4 

5 

6                    7 

8 

9 

8.8 

28.1 

39.9 

3-3.5 

49.7 

.704 

.65-3 

9 

8.5 

43.2 

55.9 

57.2 

88.8 

.751 

.048 

9 

8- 

60.9 

75.2 

77.5 

104.6 

.811 

742 

9 

7. 

94.2 

90.7 

105.4 

.... 

.8T.O 

9 

6. 

96!6 

106.1 

91.3 

105.6 

.849 

.667 

9 

1. 

95.6 

108.5 

91.3 

105.8 

.880 

.866 

7 

6.8 

23.0 

24.7 

80.9 

46.7 

.930 

.6R2 

7 

6.5 

37.9 

.... 

50.7 

•17.0 

656 

7 

6. 

53.9 

61.2 

65.5 

83.4 

.i-80 

785 

7 

5. 

68.3 

78.6 

^2.0 

62.8 

.m 

879 

4. 

73.0 

65.9 

.... 

.£51 

7 

1. 

74.8 

84.8 

ri2!7 

83.1 

.862 

873 

5 

4.8 

21.6 

20.9 

35.5 

590 

5 

4.5 

3i'.9 

.... 

39.3 

59.9 

655 

5 

4. 

42.5 

4<M 

50.5 

61.0 

.861 

829 

5 

3. 

52.2 

60.0 

51.5 

.876 

5 

1. 

54.2 

63.0 

51.2 

60.6 

.864 

849 

3 

2.8 

15.8 

17.8 

29  9 

.595 

3 

2.5 

24.*6 

29.2 

29.2 

36.8 

'.642 

794 

3 

2. 

35.8 

31.2 

.... 

.... 

3 

1. 

32!7 

38.3 

32.1 

37.0 

.652 

868 

2 

1.8 

12.2 

14.1 

21.4 

660 

2 

1.5 

11A 

21.4 

19.9 

24.2 

.'sis 

825 

2 

1. 

22.7 

21.9 

20.6 

24.1 

.911 

855 

2 

.118 

.... 

24.9 

.... 

.... 

.... 

... 

242 


STEAM  TURBINES 


and  tests  in  which  the  final  pressure  is  kept  constant  and  the 
initial  pressure  varied  are  not  as  satisfactory. 

The  two  tables  herewith  were  compiled  from  the  results  of 
about  300  tests  recorded  in  the  article  mentioned.  Table  VIII 
is  arranged  to  show  the  weights  of  steam  discharged  by  each 
nozzle  for  different  initial  pressures  and  pressure  differences. 
The  first  six  columns  are  from  the  original  data  and  the  last  two 

TABLE  IX. 
COMPARISON  OF  FLOW  THROUGH  CONVERGING  AND  DIVERGING  NOZZLES. 


Flow  in  Ib.  per  hour. 

Initial 
pressure 
Ib.  per 
sq.  in. 
absolute. 

Final  pres- 
sure Ib. 
per  sq.  in. 
absolute. 

Ratio  of 
final  to 
initial 
pressure 
(item  2-5- 
item  1.) 

Actual. 

Ratio  of 
flow  thro' 
diverging 
nozzle  to 
flow  thro' 
converging 
nozzle. 

Flow  in  Ib. 
per  hour 
per  sq.  inch 
area  at 
throat  of 
diverging 
nozzle. 

Converging 
nozzle  No.  2 
0.213  inches 
diameter  at 
throat. 

Diverging 
nozzle  No.  < 
0.213  inches 
diameter  at 
throat. 

1 

2 

3 

4 

5 

6 

7 

128 

125.2 

.979 

87.69 

109.2 

1.246 

3070 

128 

120.9 

.944 

123.1 

195.2 

1.585 

5500 

128 

113.8 

.907 

230.0 

1.390 

6440 

128 

99.6 

.776 

207  .'2 

232.0 

1.120 

6520 

128 

85.4 

.666 

233.4 

232.4 

.996 

6521 

128 

74.0=58$ 
71.2 

of  initial  pr 
.555 

essure. 
238.7 

128 

14.22 

.111 

238.7 

233  '8 

*!975 

'6521 

99.6 

96.7 

.972 

54.34 

102.7 

1.890 

2885 

99.6 

92.5 

.928 

169.5 



4775 

99.6 

85.4 

.857 

me' 

183.4 

i'363 

5150 

99.6 

71.2 

.714 

172.9 

182.2 

1.053 

5130 

57.7=58$ 

of  initial  pr 

essure. 

99.6 

56.9 

.57 

189.0 

99.6 

14.22 

.143 

186.6 

182.5 

''977 

5134 

71.2 

68.3 

.960 

47.52 

78.1 

1.642 

2200 

71.2 

64.0 

.910 



131.8 



3710 

71.2 

56.9 

.800 

108*7' 

134.2 

i'235 

3790 

71.2 

42.7 

.599 

132.0 



.... 

41.2=58$ 

of  initial  pr 

essure. 

71.2 

14.2-2 

.200 

138.6 

133.3 

.962 

3750 

42.7 

39.8 

.934 

34.76 

65.3 

1.890 

1852 

42.7 

35.6 

.933 

64.20 

80.9 

1.261 

2275 

42.7 

28.4 

.666 

78.80 

.... 



.... 

.... 

24.8=58$ 

of  initial  pr 

essure. 

42.7 

14.22 

.333 

84.30 

81.4 

.965 

2290 

28.4 

25.6 

.910 

26.8 

47.1 

1.757 

1325 

28.4 

21.3 

.751 

47.1 

53.2 

1.130 

1510 

16.5=58$ 

of  initial  pr 

essure. 

28.'4 

14.22 

.500 

54.7 

53.0 

.969 

1492 

28.4 

1.69 

.601 

54.8 





EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


243 


columns  were  calculated  by  the  author  to  facilitate  comparisons. 

Table  IX  contains  certain  data  and  results  changed  into  English 
units  and  is  arranged  for  the  purpose  of  comparing  the  rate  of 
flow  from  the  straight  and  diverging  nozzles,  Nos.  2  and  4,  with 
rounded  inlets. 

Investigations  of  Dr.  C.  E.  Lucke. — In  a  paper  presented  before 
the  American  Society  of  Mechanical  Engineers  in  January,  1905, 
Dr.  Lucke  of  Columbia  University  records  the  results  of  ex- 
periments showing  the  pressures  and  temperatures  of  steam  when 
flowing  through  a  diverging  nozzle  such  as  is  used  in  a  De  Laval 


Distance  from  Throat  in  Inches. 


1.18          0.68  0.18  0       0.32 

! 


Fig.  12. 


turbine.  While  no  conclusions  can  be  drawn  from  these  ex- 
periments at  present,  owing  to  their  incomplete  state,  the  results 
indicate  that  the  present  method  of  converting  the  heat  of  steam 


244  STEAM  TURBINES 

into  kinetic  energy  in  a  nozzle  may  be  less  efficient  than  supposed, 
and  that  the  commonly-accepted  theory  of  the  flow  of  steam 
through  orifices  does  not  conform  to  practice.  In  any  case  the 
tests  show  that  further  investigation  is  needed  before  we  can 
assert  that  the  passages  of  a  steam  turbine  have  been  designed 
along  the  most  successful  lines  possible. 

In  Fig.  13  is  a  sketch  giving  the  dimensions  of  the  nozzle. 
The  pressures  at  the  different  points  within  the  nozzle  were 
determined  in  the  usual  manner  by  means  of  searching  tubes. 
These  were  made  in  three  different  ways — one  having  an  opening 
opposed  to  the  current,  one  with  the  orifice  opening  in  the 
direction  of  the  current,  and  the  third  with  the  opening  in  the 
side  of  the  tube.  As  a  matter  of  course  these  three  tubes  gave 
widely  different  results,  and  it  seems  better,  for  the  sake  of 
comparison  with  other  experiments,  if  for  no  other  reason,  to  use 
the  results  obtained  with  the  latter  form  of  tube,  since  this  is 
the  type  employed  by  other  experimenters. 

In  the  diagram,  Fig.  12,  are  four  curves  showing  the  variation 
of  pressure  at  different  points  within  the  tube,  starting  at  initial 
pressures  of  94.72,  84.72,  74.72,  and  54.72  pounds  absolute,  respec- 
tively. The  nozzle  discharged  against  atmospheric  pressure  and 
was  evidently  designed  for  a  greater  pressure  range  of  steam  than 
was  used.  It  will  be  noted  that  in  each  case  the  steam  reaches  at- 
mospheric pressure  midway  between  the  inlet  and  outlet  sections  of 
the  diverging  portions  of  the  nozzle,  after  which  the  effect  is  to 
cause  what  we  have  previously  called  "over-expansion"  of  the 
steam,  the  pressure  dropping  to  10  pounds  absolute  and  then 
rising  again  to  atmospheric  pressure  as  the  outlet  of  the  nozzle  is 
reached. 

The  theoretical  critical  point  of  58  per  cent  of  the  initial  pres- 
sure is  not  reached  until  a  point  about  .12  inch  from  the  throat 
is  reached,  represented  by  the  line  xy.  The  points  in  the  different 
curves  at  which  the  pressure  is  58  per  cent  of  the  initial  pressure 
fall  very  nearly  in  the  same  vertical  line.  The  actual  throat  pres- 
sures at  A,  B,  C,  and  D  are  much  higher  than  this. 

Temperature  Determinations. — In  Fig.  13  is  a  diagram  repre- 
senting temperatures  at  different  points  in  the  nozzle  for  steam 
having  an  initial  pressure  of  84.72  pounds.  The  measured  temper- 


EXPERIMENTS  ON  THE  FLOW  OF  STEAM 


245 


attires  are  represented  by  line  A.  The  temperatures  were  deter- 
mined by  means  of  a  thermo-couple  of  thin  nickel  copper  wire 
which  was  stretched  through  the  nozzle.  It  will  be  seen  from  the 
diagram  and  also  from  Table  X.  that  the  steam  was  superheated 


Ini 

1;i  1  Steam 

Pressure  8 

1.72  Ibs.  Al 

s. 

\ 

jlts^ 

-^ 

•5 

« 

®\- 

If- 

1.85"—  ' 

—  H  >i<- 

—.§£  —  > 

\ 

330 

\^Cur 

re  A  of  Me 

asured  Ter 

iperatures 

K 

-300 
-290 
-280 
-270 
-260 
-250 
-240 

h 

J2 

\\ 

1 

\ 

V 

£ 
hi) 

\ 

8 

\ 

1 

\ 

S 

\ 

f  Curve  B 

showing 

1 

3 

IS 

^2 

Degrees  S 

iperheat. 

H 

jq 

y 

/ 

\ 

\ 

^X^ 

\ 

^J\ 

JiU 

/ 

^^ 

^A 

fc 

-200 
1 

0 

Diiptane 

v>  from  T 

liroat  in 

Inches. 

18         0.68           0.18    6      0.32           0.&2           1.32          1.82       2.32 

1 

Fig.  13. 

throughout  the  entire  length  of  the  nozzle.  Because  of  throt- 
tling at  the  point  of  approach,  the  steam  was  superheated  slightly 
at  the  start  and  remained  so  until  it  discharged.  Line  B  shows 
the  amount  of  superheating  at  different  points,  and  indicates 
that  the  superheat  increased  to  about  the  middle  of  the  diverging 
part  of  the  nozzle  and  then  gradually  decreased. 

In  the  last  column  of  Table  ~K.  is  given  the  number  of  thermal 
units  in  the  steam  just  before  it  enters  the  nozzle  and  at  the  point 


246 


STEAM  TURBINES 


TABLE  X. 
TEMPERATURES  IN  AN  EXPANDING  NOZZLE. 


Distance 
from 
Throat, 
Inches. 

Absolute 
Pressure, 
Ibs.  sq.  in. 

Corre- 
sponding 
Tempera- 
ture 
Saturated 
Steam. 

Measured 
Tempera- 
ture. 

Degrees 
Super- 
heat. 

B.  T.  U.  in 
Steam  =\ 

+  .48  X 

(ts-t) 

—.18 

84.72 

315.79 

330.7 

14.91 

1185.4 

.12 

49.72 

280.5 

310.2 

29.7 

.32 

27.22 

244.8 

285.7 

40.9 

.52 

16.49 

217.9 

259.7 

41.8 

.82 

13.35 

207.2 

241.7 

34.5 

1.32 

9.96 

193.1 

228.2 

35.1 

1.82 

15. 

213.03 

234.7 

21.67 

1157.3 

of  discharge..  These  were  calculated  by  taking  the  total  heat 
from  a  steam  table  and  adding  the  amount  of  heat  due  to  the 
superheat,  found  by  multiplying  the  number  of  degrees  of  super- 
heat by  the  specific  heat  at  atmospheric  pressure,  which  is  .48. 
The  difference  between  the  heat  present  at  the  beginning  and 
end  represents  the  heat  energy  of  the  steam  converted  into  work, 
as  equal  to  28.1  thermal  units. 

If  the  steam  had  been  saturated  at  the  start,  however,  and,  as 
is  generally  assumed,  the  expansion  had  been  adiabatic,  about  10 
per  cent  of  the  steam  would  have  been  condensed  and  its  heat 
converted  into  work;  and  furthermore,  the  heat  represented  by 
the  conversion  would  be  nearly  five  times  as  much  as  was  actually 
converted  into  kinetic  energy. 

The  questions  to  be  answered  are:  Why  was  so  little  heat 
converted  into  work?  and,  Why  did  the  steam  superheat  instead 
of  condense  within  the  nozzle?  Further  experiments  are  to  be 
conducted  along  these  lines,  which  should  be  productive  of  much 
valuable  information. 


CHAPTER  XII 

STEAM  AND  ITS  PROPERTIES. 

It  is  the  purpose  of  this  chapter  to  place  before  the  reader  the 
definitions,  data,  etc.,  which  form  the  basis  of  steam  calculations 
in  steam  turbine  work.  Such  calculations  pertain  mainly  to  the 
flow  of  steam  and  to  steam  nozzle  design,  which  are  treated  in  the 
next  chapter. 

The  following  is  a  list  of  the  symbols  and  their  meaning  which 
appear  in  this  and  the  succeeding  chapter. 

NOTATION  (ENGLISH  SYSTEM  ASSUMED). 

/>=absolute  pressure,  pounds  per  square  inch. 
P=absolute  pressure,  pounds  per  square  foot. 

f=temperature,  degrees  F. 
T=absolute  temperature. 

/^mechanical  equivalent  of  heat  (778  foot-pounds). 
g=:heat  of  the  liquid. 
r=latent  heat  of  vaporization. 

X  (Greek  letter  lambda)  =total  heat  of  dry,  saturated  steam. 

6  (Greek  letter  theta)=entropy  of  water. 

0  (Greek  letter  phi)i=entropy  of  steam. 

c=specific  heat  at  constant  pressure. 
z;=specific  volume  (general  expression), 
^—specific  volume  of  dry,  saturated  steam. 

<r(  Greek  letter  sigma)=:specific  volume  of  water=0.016  cubic  foot. 
F=velocity  in  feet  per  second, 
^^acceleration  due  to  gravity=32.2. 

V* 

-^-r^kinetic  energy  of  a  jet  in  foot-pounds. 

*=:per  cent  dry  steam  in  a  mixture  of  steam  and  water, 
in  square  feet, 
in  square  inches. 
J^/— weight  in  pounds. 

Temperature. — The  reference  points  of  temperature  are  the 
melting  point  of  ice  and  the  boiling  point  of  water  at  average 
atmospheric  pressure.  These  points  are  0°  and  100°  on  the  centi- 
grade scale  and  32°  and  212°  on  the  Fahrenheit  scale.  One  hun- 
dred degrees  C.  correspond  to  180  degrees  F.  Hence,  to  convert 
centigrade  units  to  Fahrenheit  units,  multiply  by  %  and  add  32. 


248  STEAM  TURBINES 

To  convert  Fahrenheit  to  centigrade,  subtract  32  and  multiply 

by  %• 

Example.— 140°  C.:=140X9/5+32=2840  F.  Again,  5°  F.  (or 
27°  below  freezing)  =  (5— 32)X%=— WX%=— 15°  C.,orl5° 
below  zero,  C. 

Absolute  Temperature. — The  absolute  zero  of  temperature,  at 
which  heat  is  supposed  to  be  entirely  absent,  is  theoretically 
492.7  degrees  F.  below  the  freezing  point  of  water.  The  demon- 
stration of  this  given  in  treatises  upon  heat  is  based  upon  the  law 
of  the  heat  expansion  of  gases,  under  the  assumption  that  the  law 
holds  at  the  extremely  low  temperatures.  This  assumption  is  in 
error,  but  to  what  extent  is  not  known.  Letting  T=  absolute  tem- 
perature and  ^^temperature  on  the  ordinary  scale, 

7'=H-460.7  Fahrenheit,  and 
T=H-273.7  centigrade. 

In  steam  calculations  absolute  temperatures  are  not  to  be  used 
unless  it  is  so  stated.  The  temperatures  of  the  steam  table  are  the 
ordinary  temperatures. 

Pressure. — The  average  atmospheric  pressure  is  taken  in  this 
country  to  be  14.7  pounds  per  square  inch.  On  the  continent  of 
Europe  pressures  are  usually  measured  in  atmospheres,  and  one 
atmosphere  is  taken  to  be  equal  to  a  pressure  of  one  kilogram  per 
square  centimeter,  which  is  equivalent  to  14.22  pounds  per  square 
inch,  instead  of  14.7  pounds. 

Gauge  pressure  is  the  pressure  denoted  by  a  steam  gauge  and  is 
measured  above  the  pressure  of  the  atmosphere;  that  is,  it  does 
not  include  the  pressure  of  the  atmosphere. 

Absolute  pressure  is  equal  to  gauge  pressure  plus  atmospheric 
pressure,  usually  taken  at  14.7  pounds  per  square  inch.  The  exact 
absolute  pressure  can  be  determined  only  by  the  use  of  the 
barometer,  which  gives  the  pressure  of  the  atmosphere  in  inches 
of  mercury.  One  cubic  inch  of  mercury  weighs  0.49  pound  at 
60  degrees  F.  and  30  inches  of  mercury  are  therefore  equivalent 
to  14.7  pounds  pressure  at  this  temperature. 

In  steam  calculations  absolute  pressures  are  to  be  used  instead 
of  gauge  pressures,  unless  otherwise  stated. 

Vacuum  is  measured  in  inches  of  mercury  or  millimeters  of 


STEAM  AND  ITS  PROPERTIES  249 

mercury.  The  vacuum  gauge  used  in  commercial  work  shows  the 
height  in  inches  of  a  column  of  mercury  that  the  pressure  of  the 
atmosphere  will  support  against  the  pressure  that  is  being 
measured. 

Example. — If  a  vacuum  gauge  attached  to  a  condenser  reads 
26  inches,  and  the  barometer  stands  at  30  inches,  what  is  the  abso- 
lute pressure  in  the  condenser?  Since  the  barometer  stands  at 
30  inches,  the  atmospheric  pressure  is  30X0.49=14.7  pounds  per 
square  inch.  30 — 26=4;  and  the  pressure  in  condenser= 
%0X14.7=1.96  pound  per  square  inch. 

Heat  Unit,  or  Thermal  Unit. — In  the  English  system  heat  is 
measured  in  British  thermal  units  (B.  T.  U.).  A  British  thermal 
unit  is  the  amount  of  heat  necessary  to  raise  one  pound  of  water 
from  62  degrees  F.  to  63  degrees  F.  In  the  French  system  the 
calorie  is  the  unit,  equal  to  3.968  B.  T.  U. 

Mechanical  Equivalent  of  Heat. — This  is  the  number  of  units 
of  mechanical  work  to  which  one  unit  of  heat  is  equivalent. 

1  B.  T.  U.=T78  foot-pounds. 

1  calorie=426.9  meter-kilograms. 

Specific  Heat  is  the  number  of  thermal  units  required  to  raise 
unit  weight  of  a  substance  one  degree  temperature.  The  specific 
heat  of  water  is  unity  at  temperatures  ranging  from  59  to  68  and 
from  104  to  113  degrees  F.,  and  approximately  unity  at  other 
temperatures.  The  specific  heat  of  nearly  all  other  substances  is 
less  than  one. 

The  foregoing  is  properly  the  definition  for  true  specific  heat. 
Sometimes,  when  a  substance  is  raised  through  several  degrees 
temperature,  it  is  necessary  to  find  the  mean  specific  heat  between 
these  limits  of  temperature.  This  is  the  average  number  of 
thermal  units  per  degree  required  to  raise  a  unit  weight  of  a  sub- 
stance from  one  temperature  to  any  other  given  temperature. 

Specific  Volume. — In  problems  relating  to  the  expansion  of 
gases,  it  is  convenient  to  deal  with  unit  weights  of  the  substance 
and  to  consider  the  volume  occupied  by  unit  weight.  This  is 
called  specific  volume.  In  the  English  system  it  is  the  cubic  feet 
occupied  by  one  pound  and  in  the  metric  system  the  cubic  meters 
occupied  by  one  kilogram. 


250  STEAM  TURBINES 

Specific  Pressure. — When  specific  volumes  enter  into  an  exam- 
ple we  have  to  deal  with  cubic  feet  instead  of  cubic  inches  and 
with  cubic  meters  instead  of  cubic  centimeters.  Pressures  used 
in  carrying  through  the  calculation  must  therefore  be  expressed  in 
pounds  per  square  foot  or.  in  kilograms  per  square  meter,  accord- 
ing to  which  system  is  being  used.  These  are  called  specific 
pressures.  Errors  frequently  creep  into  calculations  by  failure  to 
note  when  specific  pressures  should  be  used. 

Saturated  Steam  is  steam  generated  in  contact  with  water  and 
the  temperature  of  which  always  corresponds  with  the  pressure. 

Superheated  Steam  is  steam  heated  to  a  temperature  higher 
than  that  corresponding  to  the  pressure,  as  in  saturated  steam. 
The  superheating  is  produced  by  applying  heat  directly  to  the 
steam  itself,  instead  of  to  the  water  from  which  it  is  generated, 
and  if  the  superheating  is  to  be  carried  far,  the  steam  must  be  in 
a  separate  chamber,  and  not  in  contact  with  water. 

Steam  Tables. — Tables  of  the  properties  of  saturated  steam, 
used  to  facilitate  steam  calculations,  contain  columns  of  figures 
giving  certain  important  properties  of  steam.  The  following  are 
the  most  important  headings  for  steam  turbine  calculations. 
Columns  of  the  table  giving  heat  units  refer  to  the  number  of  heat 
units  in  one  pound  of  water  or  steam,  as  the  case  may  be. 

1.  Absolute  pressure,  pounds  per  square  inch. 

2.  Temperature  of  the  boiling  point  corresponding  to  pressures 
of  column  1. 

3.  Heat  of  the  liquid,  from  32°  F. 

4.  Latent  heat  of  vaporization. 

5.  Total  heat  in  the  steam  and  water,  from  32°  F. 

6.  Entropy  of  water. 

7.  Entropy  of  steam. 

8.  Specific  volume — cubic  feet  per  pound. 

9.  Density — weight  of  a  cubic  foot  in  pounds. 

Of  these,  numbers  6  and  7  will  be  explained  later,  while  num- 
bers 3,  4,  and  5  can  be  best  illustrated  by  considering  the  several 
steps  involved  in  the  generation  of  steam. 

The  Generation  of  Steam. — In  the  operation  of  a  steam  boiler 
the  pressure  is  so  nearly  constant  that  it  may  be  assumed  to  be  so 
during  the  evaporation  of  each  individual  pound  of  water.  When 


STEAM  AND  ITS  PROPERTIES  251 

iteam  is  generated  under  constant  pressure,  the  process  may  be 
divided  into  several  different  steps,  as  follows : — 

First  Step. — Heating  the  water  from  32°  F.,  to  the  temperature 
of  vaporization  (equal  to  the  temperature  of  the  steam  in  the 
boiler) .  The  number  of  heat  units  required  is  called  the  "heat  of 
the  liquid,"  represented  by  q. 

Second  Step. — Changing  the  water  into  steam,  during  which 
process  the  temperature  does  not  change.  The  heat  goes  first  to 
break  up  and  separate  the  particles  of  water  instead  of  to  raise  the 
temperature,  as  in  the  first  step ;  and  second,  to  increase  the  vol- 
ume from  that  occupied  by  the  water  to  that  occupied  by  the 
steam.  The  number  of  heat  units  required  for  the  second  step  is 
called  the  "latent  heat  of  vaporization,"  or  "latent  heat,"  simply, 
represented  by  r.  The  total  heat  required  to  generate  steam,  in- 
cluding heating  the  water,  is  called  the  "total  heat  of  the  steam," 
and  comprises  the  total  heat  energy  of  the  steam.  (Represented 
by  X.) 

Third  Step. — If  superheated  steam  were  produced,  there  would 
be  a  third  step  consisting  in  raising  the  temperature  from  that  of 
saturated  to  that  of  the  superheated  steam. 

Heat  Required  to  Raise  Temperature  of  Water. — Since  the  spe- 
cific heat  of  water  is  approximately  1  at  all  ordinary  temperatures, 
the  number  of  heat  units  required  to  raise  the  temperature  of  one 
pound  of  water  from  t1  to  t2  can  be  approximately  calculated  by 
subtracting  /x  from  t2,  or 

Heat  required=£2 — t^  (1) 

Also  the  heat  q  of  the  liquid  can  be  found  approximately  by  mak- 
ing ^=32  degrees ;  thus, 

q=t—32  (2) 

where  t  is  the  temperature  of  the  water. 

Heat  Contained  in  Wet  Steam. — Saturated  steam  condenses 
rapidly  when  its  heat  is  converted  into  mechanical  work,  or  when 
it  comes  into  contact  with  a  colder  body,  and  in  steam  calculations 
we  have  generally  to  deal  with  a  mixture  of  steam  and  water 
instead  of  with  dry  steam. 

The  quantity  of  steam  present  in  a  mixture  of  steam  and  water 
is  expressed  as  a  certain  number  of  hundredths,  x,  of  the  total 


252  STEAM  TURBINES 

weight  of  the  mixture.  One  pound  of  the  mixture  contains  a  por- 
tion x  of  steam  and  (1 — x)  of  water.  The  total  heat  of  the  mix- 
ture is  equal  to  the  heat  required  to  raise  one  pound  of  water  to 
the  given  pressure,  plus  the  heat  required  to  evaporate  the  part  x 
of  the  water  into  steam ;  or,  q-\-xr.  Hence  total  heat  of  one 
pound  of  wet  steam =  xr-\-q.  (3) 

Specific  Volume  of  Wet  Steam. — : 

Let  j— specific  volume  of  dry  steam. 
v= specific  volume  of  wet  steam. 

<r=specific    volume    of    water=l/62.4=0.016,    since    the 
weight  of  one  cubic  foot  of  water  is  62.4  pounds. 
Then,  v=xs+(l—x)<r.  (4) 

The  last  term,  (1 — JT)J>,  is  so  small  that  it  can  usually  be 
omitted.  At  high  pressures,  say  at  200  pounds,  specific  volumes 
have  low  values  and  the  effect  of  the  term  (1 — x)v  is  proportion- 
ately great.  At  200  pounds  pressure,  however,  its  omission  would 
cause  an  error  of  less  than  a  tenth  of  one  per  cent  for  each  10  per 
cent  of  moisture  present,  which  is  less  than  probable  errors  in  the 
steam  table.  We  may,  therefore,  use  for  specific  volume  of  wet 
steam,  v=.rs.  (5) 

Total  heat  of  Superheated  Steam. — To  determine  this,  we  have, 

Total  heat=A+c  (ts—t),  (6) 

where 

A=total  heat  of  dry  saturated  steam  at  the  given  pressure. 
/s=temperature  of  the  superheated  steam, 
/^temperature  of  saturated  steam  at  the  given  pressure. 
c= specific   heat   of   superheated   steam  at   constant  pressure. 
Values  of  this  will  shortly  be  discussed. 

Example. — Find  the  total  heat  of  superheated  steam  at  450  de- 
grees F.  and  100  pounds  absolute  pressure,  assuming  c=0.55. 
From  the  steam  tables,  X  =1,181.9  and  f=327.3.    Hence, 
Total  heat=l,181.9+0.55  (450—327.3). 
=1,181.9+67.5=1,249.4. 

Adiabatic  Expansion. — When  steam  expands  without  receiving 
or  giving  up  heat  it  is  said  to  expand  adiabatically.  Steam  flow- 
ing through  a  correctly  proportioned  nozzle  flows  adiabatically,  or 


STEAM  AND  ITS  PROPERTIES 


253 


nearly  so,  because  its  passage  is  so  rapid  that  little  or  no  heat  can 
be  transmitted  to  it  from  any  external  source,  nor  can  much  heat 
be  lost  through  radiation  or  otherwise. 

Temperature-Entropy  Diagram. 

The  graphical  method  of  representing  mechanical  work  is  by 
means  of  the  pressure-volume  diagram,  like  the  indicator  card  of 
a  steam  engine.  In  laying  out  such  a  diagram  we  use  two  co- 


Fig.    1.     Pressure- Volume    Diagram. 

ordinates,  OX  and  OY,  Fig.  1,  drawn  through  point  O,  the  zero 
of  volume  and  pressure.  Pointsmen  the  diagram  are  determined 
by  measuring  volumes  from  OX  and  pressures  from  OY.  The 
area  of  the  diagram  represents  mechanical  work. 

Heat  Diagram. — Similarly,  heat  energy  may  be  represented  by 
the  area  of  a  diagram  constructed  with  vertical  ordinates  of  abso- 
lute temperature  and  horizontal  dimensions  obtained  by  dividing 
the  number  of  heat  units  added  or  subtracted  during  any  change 
by  the  absolute  temperature  during  that  change.  The  horizontal 
distance  of  any  point  from  the  vertical  axis  OX  of  the  heat  dia- 
gram is  called  its  Entropy,  just  as  this  distance  on  the  work  dia- 
gram represents  volume.  This  term  "entropy"  gives  the  name 
Temperature-Entropy  Diagram  to  the  heat  diagram. 

The  Analogy  between  the  Work  Diagram  and  the  Heat  Dia- 
gram may  be  further  explained  by  selecting  some  one  part  of  the 
diagram,  Fig.  1,  as  the  expansion  line  be,  which  is  reproduced  in 
Fig.  2.  The  area  abed  under  this  curve  represents  the  work  done 


254 


STEAM  TURBINES 


during  the  change  brought  about  by  the  expansion  from  b  to  c. 
The  mean  pressure  during  this  change  is  h.  The  volume  at  the 
start  is  represented  by  distance  Oa;  the  volume  at  the  end  by  0 d: 
and  the  change  in  volume  due  to  the  expansion  by  ad.* 


x          1 

h 

k 

7 

%j 

c 

j 

/      " 

1 

y\ 

2 

V/////// 

a             a 
Entropy  Units 

/change  of\ 
[  volume  in  \ 
Vcubic  feet./ 


Fig.  3. 

Now  the  area  a&cJ=product  of  width  by  mean  height,  or  ad 
byh. 

Hence,  for  the  pressure- volume  diagram : — 

/Work  done  in  foot-pounds\  /™ean  Prf  sure\ 

...  \     =    I  during  change  ] 

\       during  any  change       )  Vinlb.  per  sq.  ft./ 

In  Fig.  3  is  the  corresponding  temperature-entropy  diagram,  in 
which  be  is  the  expansion  curve  showing  a  change,  both  in  abso- 
lute temperature  and  entropy.  The  area  abed  under  this  curve 
represents  the  heat  units  given  up  by  the  working  fluid  during  the 
expansion.  The  mean  absolute  temperature  during  the  expansion 
is  h.  The  entropy  at  the  start  is  shown  by  the  distance  Oa;  the 
entropy  at  the  end  by  Od;  and  the  change  in  entropy,  due  to  the 
expansion,  by  ad.  It  is  to  be  noted  that  the  entropy  is  measured 
from  the  ordinate  OX  and  that  ad  is  not  the  entropy  of  point  d, 
but  that  it  is  the  change  in  entropy  during  the  change  in  the  condi- 
tion of  the  substance. 


*In  the  English  system  the  work  is  in  foot-pounds,  the  pressures  are  in  pounds  per 
square  foot  and  the  volumes  in  cubic  feet.  This  is  really  what  is  represented  by  the 
indicator  diagram,  although  the  pressures  are  always  measured  in  pounds  per  square 
inch  instead  of  pounds  per  square  foot.  This  is  balanced,  however,  by  taking  the 
area  of  the  piston  in  square  inches  instead  of  square  feet,  so  the  final  result  is  the 
same. 


STEAM  AND  ITS  PROPERTIES  255 

Now,  as  before,  area  abcd=product  of  width  by  mean  height, 
or  ady^h,  and  we  therefore  have  the  following  statement  for  the 
heat  diagram,  analogous  to  the  above  statement  for  the  work 
diagram  :  — 

/Number  of  heat  units  added  or\  /mean  absolute\          /change  of\ 

/subtracted  during  any  change  I   =   {  temperature   I    X  I       ,  ) 

Vper  pound  of  the  working  fluid'          Vduring  change'          \  el 

From  this  we  have, 

Change  of  Entropy  = 

Heat  units  added  or  subtracted  during  change 

Mean  absolute  temperature  during  change. 
This  we  will  take  for  our  definition  of  entropy. 

Entropy  of  Water.  —  When  water  is  heated  its  temperature  and 
entropy  both  vary,  which  makes  it  necessary  to  use  the  calculus  to 
obtain  the  equation  for  change  of  entropy. 

Assume  heat  to  be  added  to  one  pound  of  water,  raising  the 
temperature  a  small  amount,  dT. 

If  c  is  the  specific  heat  of  water,  the  heat  added  will  be  cdT. 

Let  d6  be  the  change  in  entropy  corresponding  to  the  change 
dT  in  temperature. 

Now  if  these  quantities  dT  and  dO  be  taken  indefinitely  small, 
the  temperature  during  the  change  may  be  assumed  constant  and 
equal  to  T.  Hence,  by  the  definition  of  entropy, 

cdT 


Let  us  assume  the  water  to  be  heated  from  the  temperature  7\ 
to  temperature  T2,  and  the  entropy  to  increase  from  #1  to  02.  Then, 
by  integrating  between  the  limits  7\  and  T2,  according  to  the  prin- 
ciples of  the  calculus,  we  get  for  the  change  of  entropy, 

4r-*i=*-log.5  (8) 

7\ 

Entropy  is  reckoned  from  32  degrees  F.,  or  492.7  degrees  abso- 
lute, the  same  as  the  other  heat  properties  of  the  steam  table,  and 
if  7\  be  taken  equal  to  492.7,  0L  becomes  zero  and  we  have  for  the 
entropy  of  water, 


256  STEAM  TURBINES 

T 


(.u) 


It  answers  practical  requirements  to  take  the  specific  heat,  c,  in 
the  above  formulas  equal  to  1. 

Entropy  of  Saturated  Steam.  —  The  total  heat  of  steam  is  con- 
sidered made  up  of  two  parts,  the  heat  q  required  to  raise  the  tem- 
perature of  the  water  from  the  freezing  point  to  the  temperature 
of  vaporization,  and  the  latent  heat  r  required  to  convert  the  water 
into  steam.  In  like  manner  the  entropy  of  steam  consists  of  two 
parts,  the  entropy  of  water  at  the  temperature  of  vaporization  and 
the  change,  or  increase  of  entropy  that  occurs  when  the  water  is 
changed  into  steam.  The  latter  is  sometimes  called  the  entropy  of 
vaporization,  but  more  properly  it  is  the  change  of  entropy  due  to 
vaporization. 

During  vaporization  the  temperature  remains  constant  and  the 
change  of  entropy  is  easily  calculated  by  dividing  the  latent  heat 
r  by  the  absolute  temperature  T,  or, 

r 
Change  of  entropy^— 

From  this  we  have,  for  the  entropy  of  steam,  introducing  x  to 
make  it  general  for  either  wet  or  dry  steam  (see  formula  3), 

xr 


xr 

+—  (13) 

T 

Example.  —  To  find  the  entropy  of  saturated  steam  at  100 
pounds  absolute  pressure,  we  have,  from  the  steam  tables,  />=100  ; 
r=884;  f=327.5S,  whence  T=  460.7+327.6=788.3.  Assuming 
steam  to  be  dry, 

788.3 
0=2.  3  log  - 

492.7 

='2.3x(2.  89669—  2.69258)^0.47 

884 

Hence,  <£=0.47+  --  =1.12 

788.3 


STEAM  AND  ITS  PROPERTIES  257 

Entropy  of  Superheated  Steam. — The  entropy  of  superheated 
steam  may  be  found  by  adding  to  the  entropy  of  saturated  steam 
the  change  in  entropy  due  to  superheating,  which  is  expressed  by 
the  equation, 

T* 
&- 4=2.3  clog—  (14) 

where  <£s — <£  is  the  change  in  entropy,  c  is  the  specific  heat  of 
superheated  steam,  and  7"8  and  T  are  the  temperatures  of  super- 
heated and  saturated  steam,  respectively,  at  the  given  pressure. 

Example. — If  the  steam  in  the  last  example  were  superheated 
250  degrees,  what  would  be  its  entropy,  assuming  its  specific  heat 
tpbeO.6?  Here  7=788.3  and  r8=788.3+250=l,038.3.  Hence, 

<£s_  <£=2. 3x0. 6X  (3.01632  —  2.89669) 

=2.3x0.0x0.1196=0.165 
and  <£8=1. 12+0. 165=1.285 

Temperature-Entropy  Diagram  for  Water  and  Steam. — In 
Fig.  4  the  various  heat  changes  for  water  and  steam  are  shown 
graphically.  Absolute  temperatures  are  laid  off  on  OX  and 
values  for  entropy  on  OY.  The  different  steps  in  the  process  of 
plotting  the  diagram  are  as  follows  : — 

(1)  Assume  one  pound  of  water  at  100  degrees   F.  to  be 
heated  until  its  temperature  reaches  350  degrees  F.     Its  entropy 
increases  and  the  change  in  temperature  and  entropy  is  repre- 
sented by  the  water  line  ab.     This  curve  starts  on  the  ordinate 
OX,  at  the  freezing  point,  and  other  points  on  the  curve  are  cal- 
culated by  the  aid  of  equation  (10)  ;  or  they  may  be  plotted  from 
values  of  "entropy  of  the  liquid"  given  in  steam  tables.     If  ^ 
is  the  entropy  at  point  a  and    08  at  point  b,  then   02  —  Ol  is  the 
change  in  entropy  while  the  temperature  increases  from  100  to 
350  degrees.     The  heat  added  during  the  change  is  represented 
by  area  a^abb^ 

(2)  If  the  water,  at  350  degrees,  is  vaporized,  the  temperature 
will  remain  constant  and  the  line  be  will  represent  the  change  in 
entropy.    At  point  c  the  vaporization  is  complete  and  its  distance 
from  OX  is  the  entropy  of  steam  at  350  degrees  temperature,  cal- 
culated by  equation  (12).     The  heat  required  for  this  change  is 
shown  by  the  area  b^bcc^. 


258 


.STEAM  TURBINES 


(3)  In  the  flow  of  steam  through  nozzles,  it  is  assumed  that 
the  expansion  is  adiabatic,  no  heat  being  added  or  subtracted. 
This  condition  can  be  represented  only  by  the  vertical  line  cd, 
which  has  no  area  under  it  to  indicate  heat  added  or  subtracted. 

I/ 


810.7H 


530.7- 
492.7- 


0.2        0.4       0.6 


0.8        1.0        1.2        1.4 
Scale  of  Entropy 


1.6        1.8       2.0 


Fig.    4.    Temperature-Entropy    Diagram. 

This  line  shows  that  while  the  temperature  drops  the  entropy  re- 
mains constant. 

(4)  To  complete  the  cycle,  assume  the  steam  remaining  at  the 
end  of  adiabatic  expansion  to  condense  at  constant  temperature, 
as  shown  by  the  line  da.  The  heat  given  up  during  con- 
densation is  represented  by  the  area  a^adc^.  Deducting  this  area 
from  areas  a^abb^+bjbcc^  which  show  the  heat  applied,  leaves 


STEAM  AND  ITS  PROPERTIES 


259 


the  net  area  abed,  which  represents  the  available  heat  energy  for 
doing  work  on  this  cycle. 

Line  for  Dry  Saturated  Steam. — If,  instead  of  adiabatic  ex- 
pansion, there  were  heat  enough  added  to  the  steam  while  the 
temperature  was  dropping  from  350  to  100  degrees  to  maintain 
the  steam  in  a  dry  and  saturated  state,  the  conditions  would  be 
represented  by  the  steam  line  ce.  The  heat  required  to  maintain 
this  condition  is  represented  by  c^cee^.  The  various  points  on  the 
steam  line  are  calculated  by  adding  the  entropy  due  to  vaporization 
to  the  entropy  of  water  for  the  corresponding  temperatures.  In 
other  words,  line  ce  shows  the  entropy  of  dry,  saturated  steam  for 
the  temperatures  within  its  limits. 

Line  for  Superheated  Steam. — If  at  point  c  heat  were  added  to 
superheat  the  steam  the  change  would  be  represented  by  a  curve 
like  cf,  plotted  by  the  aid  of  equation  (14).  The  heat  required  to 
superheat  is  shown  by  area  c^cff^. 

Conditions  with  Moisture  Present. — In  Fig.  5  ab  and  ce  are  the 


0--, 

7 

'492.7 L 


Fig.  5. 

water  and  steam  lines,  respectively,  of  a  temperature-entropy  dia- 
gram. Take  the  case  of  one  pound  of  water  at  temperature  T2,  as 
indicated  on  the  diagram.  Its  entropy  is  equal  to  the  distance  of 
point  m  to  the  right  of  line  OX.  If  this  water  is  all  evaporated 
into  steam,  the  conditions  will  be  represented  by  point  k  on  steam 
line  ce.  If,  however,  only  a  part  of  the  water  be  evaporated  into 


260  STEAM  TURBINES 

steam,  say  80  per  cent,  then  the  change  of  entropy  due  to  vaporiza- 
tion will  be 

xr     0.8r 


T       T 

and  the  conditions  will  be  represented  by  point  n,  which  is  80  per 
cent  of  the  distance  from  m  to  k.  By  the  principles  of  percentage, 
therefore,  the  ratio 

mn 

mk 

will  give  the  percentage,  x,  of  dry  steam  present. 

The  chief  value  of  the  temperature-entropy  diagram  for  steam 
turbine  work  lies  in  the  fact  that  it  may  be  made  to  show  the 
quantity  of  moisture  present  in  steam  by  the  method  just  indicated. 

Thus,  in  the  above  case,  suppose  the  steam  to  expand  adiabatic- 
ally  to  temperature  7\.  By  drawing  the  adiabatic  line  np,  we 
find  the  percentage  of  dry  steam  present  to  be 

of_ 

OS 

If  dry,  saturated  steam  expands  adiabatically  from  T2  to  7\, 
we  find,  by  drawing  adiabatic  line  k d  that  there  is 

od 

OS 

per  cent  of  dry  steam  present. 

If  superheated  steam  expands  adiabatically  from  temperature 
Ts  to  7\,  draw  the  adiabatic  line  fr.  At  point  /,  where  it  inter- 
sects the  steam  line,  the  steam  loses  its  superheat  and  becomes 
saturated,  while  at  temperature  7\  it  is  only 

or 

OS 

per  cent  dry.  If  it  were  desired  to  know  how  far  to  continue 
superheating  to  secure  dry,  saturated  steam  at  the  end  of  ex- 
pansion, the  superheat  curve  kf  must  be  extended  until  point  / 
comes  vertically  over  point  s,  which  would  show  the  desired  tem- 
perature Ts. 


STEAM  AND  ITS  PROPERTIES  261 

Characteristic  Equations  for  Adiabatic  Expansion. 

During  adiabatic  expansion  heat  is  neither  added  nor  abstracted 
and,  as  seen  above,  the  change  is  represented  by  a  vertical  line  on 
the  temperature-entropy  diagram.  The  entropy,  therefore,  remains 
constant  during  the  adiabatic  expansion,  and  by  expressing  this 
relation  in  the  form  of  an  equation  the  percentage  of  dry  steam 
present  at  the  end  of  adiabatic  expansion  may  be  easily  calculated. 
This  relation  is  expressed  in  the  three  equations  which  follow,  in 
which  the  letters  with  subscript  1  refer  to  the  higher  pressure  and 
those  with  subscript  2  to  the  lower  pressure. 

(1)  For  saturated  steam, 

x\r\  xiri 

-^+«i=-£+0,  (15) 

A  ^  2 

(2)  Steam    superheated    sufficiently   to   remain    superheated 
throughout  expansion, 

X-iT-i  1  ai      Xtf*  J-  s2 

+  6l  +  *.3c\oS—  =+0a  + 2.3  Hog—  (16) 

*  1  *l  *2  J-  2 

(3)  Steam  superheated  at  the  start  but  saturated  at  the  end  of 
expansion, 

X\T\  Fa    xzrz 

^r1  +  01  +  2.3flog-B  =  ^r8+02  (17) 

^1  ^1^2 

The  application  of  these  formulas  will  be  shown  in  the  follow- 
ing chapter. 

The  Specific  Heat  of  Superheated  Steam. 

Regnault's  Result. — Until  recently  the  value  universally 
adopted  for  the  specific  heat  of  superheated  steam  at  constant 
pressure  has  been  0.48,  derived  in  1840  from  the  results  of  three 
series  of  experiments  by  Regnault.  In  steam  calorimeter  work, 
where  the  temperatures  came  within  the  limits  of  Regnault's  ex- 
periments, the  value  of  0.48  is  practically  correct,  but  with  the 
high  pressures  and  temperatures  prevailing  in  power  plants  using 
superheated  steam  a  higher  value  should  be  taken. 

The  Importance   of  a   Correct   Value. — In  tests  upon  super- 


262  STEAM  TURBINES 

heaters,  or  upon  turbines  and  engines  using  superheated  steam,  the 
weight  of  the  steam  used  does  not  afford  a  fair  basis  for  estimat- 
ing the  gain  or  loss  from  superheating,  because  weight  alone 
gives  no  indication  of  the  amount  of  heat  in  superheated  steam  of 
a  given  pressure.  If  the  specific  heat  of  superheated  steam  were 
accurately  known,  however,  this  would  give  us  the  means  of  cal- 
culating the  number  of  heat  units  in  the  steam  and  the  efficiency 
of  the  apparatus  could  be  determined  on  this  basis,  by  the  method 
explained  under  "The  Thermal  Unit  Basis  of  Performance"  in 
Chapter  IX.  The  author  has  made  calculations  upon  tests  of  a  De 
Laval  turbine,  run  first  with  saturated  and  then  with  superheated 
steam.  By  first  taking  the  rate  of  water  consumption  in  pounds  as 
the  basis  for  calculating  the  gain  from  superheating,  the  gain  was 
found  to  be  8.1  per  cent.  Then,  by  taking  as  the  basis  the  heat 
units  in  the  steam,  the  gain  was  found  to  be  as  follows :  Assum- 
ing specific  heat=0.48,  gain=4.8  per  cent;  specific  heat=:0.6, 
gain—-!  per  cent;  specific  heat=0.8,  gain=2.7  per  cent.  These 
results  show  that  the  gain  from  superheating  on  the  basis  of  heat 
units  utilized  is  much  less  than  when  on  the  basis  of  pounds  of 
water  per  horse-power  per  hour;  and  that  the  higher  the  value 
assumed  for  specific  heat  the  less  the  gain  is  found  to  be.  This 
illustration  shows  the  importance  of  a  correct  value  for  the  spe- 
cific heat  of  superheated  steam  in  calculating  efficiencies. 

Results  of  Tests  to  Determine  Specific  Pleat.* — Many  experi- 
menters have  attempted  to  derive  values  for  the  specific  heat  of 
superheated  steam  at  constant  pressure  for  other  pressures  and 
temperatures  than  covered  by  the  tests  of  Regnault.  Among  the 
more  important  work  in  this  connection  is  that  of  Grindley  in 
England,  of  Greissmann,  Lorenz,  and  Messrs.  Knoblauch,  Linde 
and  Klebe  in  Germany,  and  of  Carpenter,  Jones,  Thomas,  Bur- 
goon,  and  engineers  of  the  General  Electric  Company  in  America. 
The  results  of  the  various  experimenters  are  more  or  less  contra- 
dictory and  it  is  not  yet  definitely  settled  how  the  specific  heat 
varies  in  relation  to  pressure  and  temperature  changes. 

*The  reader  who  wishes  to  investigate  the  subject  of  specific  heat  of  superheated 
steam  is  referred  to  the  Journal  of  the  Worcester  Polytechnic  Institute,  November, 
1904,  containing  an  article  by  Prof.  Sidney  A.  Reeve;  to  Power  for  August,  1904, 
containing  an  article  by  Chas.  A.  Orrok;  to  the  Stevens  Institute  Indicator  for  October, 
'  1905,  containing  an  article  by  Prof.  T.  E.  Denton;  and  to  the  paper  upon  the  "Steam 
Plant  of  the  White  Motor  Car,"  by  Carpenter,  in  the  proceedings  of  the  A.  S.  M.  E., 
December,  1906. 


STEAM  AND  ITS  PROPERTIES  263 

Greissmann's  and  Grindley's  Results.  —  An  early  review  of  the 
work  of  the  different  experimenters  was  contributed  by  Mr. 
George  A.  Orrok,  chief  draftsman  of  the  New  York  Edison  Com- 
pany, to  Power  for  August,  1904.  He  plotted  the  various  values 
and  found  those  of  Greissmann  to  be  the  most  consistent.  Taking 
these  as  a  basis  Mr.  Orrok  deduces  the  following  formula  to 
represent  them  : 

£=0.00222  4—0.116  (18) 

in  which  ts  =temperature  of  the  superheated  steam. 

The  above  formula  gives  the  instantaneous  or  true  specific  heat 
at  any  temperature.  The  mean  value  of  the  specific  heat  between 
the  points  of  saturation  and  any  degree  of  superheat  can  be  found 
by  the  formula 

<r=0.00222 


/4+^\ 

(-    -1-0.116  (19) 

\    /J    ' 


A  careful  review  of  Grindley's  work  has  been  made  by  Pro- 
fessor Reeve,  of  Lawrence  Scientific  School,  who  has  recalculated 
the  values  as  originally  given.  Grindley's  results,  as  recalculated 
by  Reeve,  and  Greissmann's  values,  as  given  by  Orrok's  formulas, 
agree  quite  closely.  It  is  to  be  noted  that  Greissmann's  work 
indicated  that  the  specific  heat  increases  with  the  temperature 
without  regard  to  what  the  pressure  is  ;  that  is,  steam  of  low 
pressure  and  high  superheat  would  have  the  same  specific  heat  as 
steam  of  high  pressure  and  low  superheat,  provided  the  tempera- 
tures were  the  same.  The  conclusion  was  drawn  by  Lorenz, 
however  (London  Engineer,  July  8,  1904),  that  specific  heat  in- 
creases with  increase  in  pressure,  but  decreases  with  increase  of 
superheat  at  any  given  pressure.  This  conclusion  is  confirmed  by 
later  experiments  which  will  be  referred  to. 

Results  of  Knoblauch,  Linde  and  Klebe.  —  In  the  Stevens  In- 
stitute Indicator  for  October,  1905,  is  an  article  by  Prof.  J.  E. 
Denton,  referred  to  in  the  last  footnote,  outlining  the  remarkable 
work  done  by  these  investigators.  Their  work  was  mainly  the 
finding  of  the  volume  of  saturated  and  superheated  steam  under 
different  pressures  and  temperatures  ;  but  from  the  equation  con- 
necting these  elements  they  were  able  to  deduce  equations  for  the 
specific  heat  of  superheated  steam,  and  also  to  explain  the  reasons 


264  STEAM  TURBINES 

for  the  incongruities  in  the  results  of  some  of  the  previous  re- 
searches. A  table  prepared  by  Professor  Denton  from  their 
formula  is  given  below : 

MEAN  SPECIFIC  HEAT  AT  CONSTANT  PRESSURE. 
(KNOBLAUCH,  LINDE,  AND  KLEBE.) 

Boiler  Boiling  Point  Range  of   Superheating. 

Pressure  Degrees  10°  C.  =    50°  C.  =    100°  C.  = 

Lb.  Sq.  In.  Absolute.  C.  50°  F.        122°  F.       212°  F. 

9948  164  0.567  0.551  0.537 

139.32  178  .597  .577  .559 

190.70  192  .634  .609  .586 

266.20  208  .686  .656  .626 

Results  at  Cornell  University. — Experiments  upon  specific  heat 
of  superheated  steam  have  been  under  way  for  over  10  years  at 
Cornell  University.  The  following  table  is  made  up  from  a 
chart  giving  results  obtained  by  Prof.  Carl  C.  Thomas  and  Mr. 
C.  E.  Burgoon  at  this  university,  and  published  by  Prof.  R.  C. 
Carpenter  in  a  paper,  "Steam  Plant  of  the  White  Motor  Car," 
read  before  the  A.  S.  M.  E.  in  December,  1906. 

SPECIFIC  HEAT  AT  CONSTANT  PRESSURE. 

(THOMAS  AND  BURGOON.) 
Pressure  Degrees  F.  Superheat. 


Lb.  Sq.  In.  Abs. 

25 

50 

100 

150 

200 

50 

.515 

.513 

.512 

.510 

.509 

100 

.549 

.544 

.54 

.536 

.530 

150 

.581 

.576 

.57 

.561 

.554 

200 

.614 

.61 

.598 

.589 

.577 

250 

.645 

.642 

.629 

.613 

.602 

Specific  Volume  of  Superheated  Steam. 

While  the  specific  volumes  of  saturated  steam  for  different 
pressures  are  to  be  found  in  the  steam  tables,  if  calculations  are  to 
be  made  requiring  the  specific  volume  of  superheated  steam,  the 
information  is  not  so  readily  obtained. 

Zeuner's  Formula. — The  formula  usually  employed  is  that  of 
Zeuner,  based  upon  the  experiments  of  Hirn.  It  is  as  follows : — 

93.5T— 971P* 
*=-      Jy-  (20) 

In  this  P  is  pressure  in  pounds  per  square  foot,  or  144 X/>- 


STEAM  AND  ITS  PROPERTIES  265 

Example. — Superheated  steam,  having  a  pressure  of  100  pounds 
absolute  and  a  temperature  of  400  degrees  F.,  or  860.7  degrees 
absolute,  has  a  specific  volume  of 

__93.5X860.7— 971(144X100)* 

144X100 
_80,475— 10,635 

14,400 
—4.8  cubic  feet. 

Schmidt's  Formula. — Another  formula  that  has  been  proposed 
is  that  of  Schmidt,  given  below,  which  closely  approximates  Hirn's 
results,  though  not  as  closely  as  Zeuner's  formula.  The  differ- 
ence between  the  results  obtained  with  the  two  formulas  is  slight, 
however. 

441.4+f 

^=0.59276—  (21) 

P 

Example. — Taking  the  same  data  as  above,  we  have, 

441.44-400 

v—  0.59276 

100 

=4.98  cubic  feet. 


CHAPTER  XIII 


CALCULATIONS  ON  THE  FXOW  OF  STEAM. 
The  Adiabatic  Flow  of  Steam. 

In  calculations  on  the  flow  of  steam  it  is  assumed  that  the 
flow  is  adiabatic  and  afterwards  allowances  are  made,  if  necessary, 
based  upon  the  results  of  actual  tests.  Under  this  condition  all  the 
available  heat  energy  of  the  steam  is  assumed  to  be  converted  into 
kinetic  energy,  without  gain  or  loss  of  heat  through  conduction, 
radiation,  friction  or  otherwise,  and  the  energy  of  the  steam  will 
remain  the  same  in  amount  at  all  steps  in  the  process,  though  it 
may  differ  in  form. 

Equation  for  the  Flow  of  Saturated  Steam. — In  Fig.  1  is  a 
cylinder  having  a  diaphragm  which  separates  it  into  two  cham- 


Fig.  1. 

bers,  A  and  B.  A  nozzle  N  is  inserted  in  the  diaphragm  and 
steam  in  chamber  A,  at  absolute  pressure  plf  flows  through  the 
nozzle  into  chamber  B.  Steam  expands  within  the  nozzle  to  the 
pressure  p2. 

Let  F=velocity  of  steam  in  feet  per  second  as  it  leaves  the 
nozzle.  Also,  let  x^  r±  and  q^  apply  to  steam  at  the  pressure  plt 
in  chamber  A,  and  x2,  r2  and  q2  to  steam  at  pressure  p2,  within 
the  nozzle.  (Notation  at  beginning  of  Chap.  XII.) 

Now,  since  the  energy  remains  constant  during  the  flow,  we 
write  expressions  for  the  energy  of  one  pound  of  steam  as  it 
approaches  the  nozzle,  and  for  one  pound  of  steam  as  it  leaves  the 
nozzle,  and  place  one  equal  to  the  other. 


CALCULATIONS  ON  THE  FLOW  OF  STEAM  267 

The  energy  of  the  steam  as  it  approaches  or  leaves  the  nozzle 
is  in  the  form  either  of  heat  energy  or  kinetic  energy,  the  latter  due 
to  its  velocity.  Usually  the  velocity  of  approach  is  so  low  that  the 
corresponding  kinetic  energy  may  be  neglected,  and  we  will  as- 
sume that  in  chamber  A  we  have  to  deal  with  heat  energy  only. 
In  turbine  work  the  initial  velocity  is  sometimes  a  considerable 
factor,  however,  and  in  such  a  case  should  be  taken  into  account. 

Heat  energy  of  one  pound  of  steam  in  chamber  A  is  x±  rl-\-q1 
heat  units,  equivalent  to  /  (x^r^q^)  foot  pounds.  (3)  Chap.  XII. 

Heat  energy  of  one  pound  of  steam  as  it  leaves  the  nozzle  is 
x2r2-\-q2  heat  units,  equivalent  to  J' (#2r2-{-q2)  f°ot  pounds. 

Kinetic  energy  of  one  pound  of  steam  as  it  leaves  the  nozzle  is, 
by  the  principles  of  mechanics, 

F2 

—  foot-pounds. 

Placing  the  energy  of  discharge  equal  to  the  energy  of  ap- 
proach, 

— +f(x,r2+q2)=J(xlr1+ql),  (1) 

«s 

which  is  the  general  equation  for  the  flow  of  saturated  steam.* 
The  mechanical  energy  of  the  jet  in  foot-pounds  is 


^  \XA  -  ^.         I          J.    J-  J    4i    /  \  / 

The  velocity  of  discharge  in  feet  per  second  is 

(3) 


=8.0»V778(*J  r—  *2  r2+q—q2)  (4) 

Value  of  x2  in  Above  Equations.  —  In  order  to  use  these  equa- 
tions, we  must  first  calculate  x2,  the  percentage  of  dry  steam  at  the 
end  of  adiabatic  expansion  in  each  case.  To  obtain  x2  use  the 

*There  should  also  be  included  as  part  of  the  energy  of  the  steam  at  the  initial 
pressure,  the  work  required  to  pump  one  pound  of  water  from  the  final  pressure  p2  to 
the  initial  pressure  pr  This  is  represented  by  the  expression 


and  is  so  small  in  amount  that  it  can  be  neglected. 


268  STEAM  TURBINES 

characteristic  equation  (15)  Chap.  XII.,  which,  when  transposed, 
becomes 

•*•!'  1 

*,-*2  + 
*,=-  (5) 


If  steam  tables  containing  values  of  0  are  not  obtainable,  it  will 
be  necessary  to  use  the  approximate  equation  (10)  Chap.  XII., 
for  finding  the  entropy  of  water  in  the  above  equation. 

Example  I.  —  Given,  dry  saturated  steam  flowing  from  a  pres- 
sure of  135  pounds  absolute  to  a  pressure  of  45  pounds  absolute. 
Calculate  the  energy  of  the  jet  and  the  velocity  of  flow,  assuming 
complete  expansion  in  the  nozzle. 

The  following  values  are  either  known  or  taken  from  the  steam 
table:— 


7\=  810.73  r2=734.99 

r^=  867.3  r2=922 

q^—  321.4  g2=243.6 

6f=          .5027  02=       .4020 

#!=  1  *2=         ? 

First,  calculate  the  percentage  of  dry  steam  at  the  end  of  the 
expansion,  from  equation  (5). 

1X867.3 
.5027—  .4020H  — 

810.73 


922 

734.99 
=.933=93.3%. 

Now  substituting  for  xz  in  equation  (2),  we  have,  for  energy  of 
the  jet, 

V2 

— =778(1X867.3— .933X922+321.4— 243.6) 

*g 

=66,032  ft.  Ib. 

and  velocity  of  discharge  is 


J/=\/64.4X66,032 
=2,062  ft.  per  sec. 


CALCULATIONS  ON  THE' FLOW  OF  STEAM  269 

When  Expansion  is  not  Complete  in  the  Nozzle. — In  working 
out  the  above  example  it  was  assumed  that  the  expansion  of  the 
steam  was  carried  to  the  terminal  pressure  of  45  pounds  within 
the  nozzle  itself  and  there  was  no  waste  energy  due  to  drop  of 
pressure  as  the  steam  left  the  nozzle.  To  accomplish  this  with  the 
pressures  given  requires  a  diverging  nozzle,  as  already  explained 
in  Chapter  I.  under  "Steam  Nozzles,"  and  in  Chapter  XI.  Sup- 
pose, however,  that  instead  of  a  diverging  nozzle  a  straight  nozzle 
with  converging  inlet  were  used.  We  have  learned  that  in  such  a 
nozzle  expansion  may  be  carried  to  a  certain  point — usually  about 
60  per  cent  of  the  higher  absolute  pressure — and  no  further,  and 
in  this  case  steam  would  expand  within  the  nozzle  to  about  80 
pounds,  which  should  be  used  for  the  lower  pressure  p2  in  the  cal- 
culations. In  any  case  where  expansion  is  not  complete  within 
the  nozzle,  care  should  be  taken  to  assume  for  p2  the  pressure  to 
zvhich  steam  expands  within  the  nozzle  itself  instead  of  the  lower 
outside  pressure. 

Example  II. — Assuming  a  straight  nozzle,  we  have  for  values 
corresponding  to  />2=80  pounds,  r2=895.6,  q2=2S!A,  02=0.452, 
T2=772.5. 

Values  corresponding  to  p±  are  given  under  Example  I. 

From  (5),  we  find  #2=0.966. 

From  (3), 

F=V64.4X778(867.3— .966X895.6+321.4— 281.4) 
=1,451.4  ft.  per  sec. 

Chart  Giving  Values  of  x. — To  assist  the  reader  in  determining 
the  percentage  of  dry  steam  at  the  end  of  adiabatic  expansion,  the 
chart  in  Fig.  2  has  been  prepared,  which  enables  the  value  of  x2 
to  be  read  directly,  without  calculation.  This  quantity  may  also 
be  easily  determined  by  the  aid  of  the  temperature-entropy  dia- 
gram, as  explained  in  connection  with  the  subject  in  Chapter  XII. 

Simplified  Formula  for  the  Flozv  of  Steam. — In  the  steam  table 
in  the  appendix  values  of  the  entropy,  <£,  of  steam  are  given,  in 
which 

<£=  — +  0 
T 


100 


92  90  88  86  84  82 

Per  Cent  of  Dry  Steam  at  End  of  Expansion 


so 


Fig.   2.     Chart   Showing   Per   Cent  of   Dry   Steam   Present  at  the   End   of  Adiabatic   Expansion. 


CALCULATIONS  ON  THE  FLOW  OF  STEAM  271 

Hence,  if,  as  is  usually  assumed  in  calculations,  the  steam  is  ini- 
tially dry,  equation  (5)  reduces  to 

?;  (*i-*,) 
*t=-  (6) 


Also,  with  steam  initially  dry,  we  may  write  A*,  the  total  heat  of 
dry  steam,  in  place  of  (&\  r^-\-q^)  in  (2),  which  gives, 

^r2-?2)  (7) 


Now  substituting  in  this  the  value  for  xz  in  (6),  we  get, 
F2 


T,(<j>i-e2)-g2]  (8) 

g 

This  gives  the  flow  directly  without  computing  ,r2. 

Example  III.  —  Assume  expansion  complete  within  the  nozzle. 

#!=    185  p2=  15 

Ai=l,196.4  T2=673.73 

<f>j=          1.5498  0=       .3143 

g2=181.8 

F2 

—=778  [1,196.4—  673.73(1.5498—  .3143)—  181.8] 
2g 

=141,757. 


=3,020  ft.  per  sec. 

Weight  of  Saturated  Steam  Flowing.  —  The  most  satisfactory, 
as  well  as  the  simplest  equations  for  the  weight  of  steam  dis- 
charged from  a  nozzle  are  Napier's  empirical  formulas  given  at 
the  beginning  of  Chapter  XL  In  the  design  of  nozzles,  however,  it 
may  sometimes  be  more  convenient  to  use  the  equations  (9  and 
10)  below,  which,  instead  of  being  based  upon  experiment,  like 
Napier's,  depend  upon  the  following  principle  :  — 

Weight  in  pounds  discharged  per  second  X  volume  in  cubic  feet 
of  one  pound  (specific  volume)  =area  of  orifice  in  square  feetX 
velocity  of  steam  in  feet  per  second. 


272  STEAM  TURBINES 

Let  J^:=weigrit  discharged  per  second  in  pounds. 
z;=specific  volume  of  the  fluid  in  the  orifice. 
A—  area,  of  orifice  in  square  feet. 
F—  velocity  of  discharge  in  feet  per  second. 
Then,  WXv=A^V 

AV 

W=-  (9) 

v 

This  formula  is  perfectly  general  and  applies  either  to  saturated 
or  superheated  steam,  or  to  any  gas  or  liquid. 

If  it  is  to  be  applied  to  saturated  steam,  then,  by  (5),  Chap.  XII., 
we  have,  v=xs,  where  x  is  the  quality  and  ^  is  the  specific  volume 
of  the  steam  in  the  orifice.  Also, 

a 

^j  —  _ 

144 
where  a  is  area  in  square  inches. 

aV 

Hence,  >  W=  -  (10) 


Example  IV.  —  What  is  the  weight  of  steam  discharged  in 
Example  II.,  assuming  the  area  of  nozzle  to  be  0.5  square  inch? 

Here  we  have,  initial  pressure  135  pounds,  and  pressure  in  noz- 
zle %0  of  this,  or  approximately  80  pounds.  The  quality  x  at  this 
lower  pressure  was  found  to  be  0.966  and  the  velocity  V  1,451 
feet  per  second.  The  specific  volume  s  corresponding  to  80 
pounds  is  5.43. 

Hence,  from  (10), 

w=    -5X1'451 

144X-966X5.43 
=0.961  pounds  per  second. 
Napier's  rule  gives  0.964  as  the  result. 

Calculations  Upon  Superheated  Steam.  i 

Equations  for  the  Flow  of  Superheated  Steam.  —  There  are  two 
cases  according  as  the  steam  is  superheated  or  saturated  when 
expansion  is  completed,  its  condition  at  that  point  depending  upon 
the  degree  of  superheat  at  the  start  and  the  degree  of  expansion 
that  takes  place.  The  calculation  of  the  velocity  or  energy  of 


CALCULATIONS  ON  THE  FLOW  OF  STEAM  273 

flow  of  superheated  steam  is  a  long  and  tedious  process,  and  until 
the  results  of  more  tests  are  available  we  are  not  sure  that  the 
calculated  results  agree  even  approximately  with  experimental 
results.  Nearly  all  tests  have  so  far  been  upon  saturated  steam. 

The  formulas  for  the  flow  of  superheated  steam  are  derived  by 
the  same  method  as  the  one  for  saturated  steam. 

The  heat  energy  of  one  pound  of  superheated  steam  is 

\  +  c(tB-t)  (6)  Chap.  XII. 

Placing  the  energy  of  discharge  equal  to  the  energy  of  ap- 
proach, and  using  letters  with  subscripts  1  and  2  to  represent  ini- 
tial and  final  conditions,  respectively,  we  have  : 

Case  I.,  when  steam  is  superheated  at  the  end, 

F2 

—  -r/[A,  +  c  (/82-  /,)1=/[A1+  c  (t«-tj\ 
"& 
or, 

V* 

—  =/[*t-VM^--*i)-*(/«-4H 

Case  II.,  when  steam  is  saturated  at  the  end, 


+  ft)  = 

"& 
or, 

F2 

—  =/[Ai  +  c  (tA  -  t,)  -  x^  -  ft]  (12) 

To  find  the  Pressure  at  which  Superheated  Steam  Loses  its 
Superheat  During  Adiabatic  Expansion.  —  Before  one  can  proceed 
to  calculate  the  velocity  of  flow  of  superheated  steam  it  must  be 
determined  whether  the  example  comes  under  Case  I.  or  Case  II.  ; 
that  is,  whether  the  steam  is  superheated  or  saturated  at  the  end  of 
the  flow.  This  is  done  by  making  jr2=l  in  characteristic  equation 
(17),  Chap.  XII.  ,  for  superheated  steam.  The  second  member  of 
this  equation  will  then  be 


which  is  the  entropy  of  dry,  saturated  steam,  and  the  equation  will 
express  the  relation  that  the  entropy  of  superheated  steam  at  a 


274  STEAM  TURBINES 

certain  pressure  and  temperature  equals  the  entropy  of  dry, 
saturated  steam  at  a  certain  pressure,  which  latter  pressure  is  to  be 
determined.  After  making  xz—\,  find  the  value  of  the  left-hand 
member  of  the  equation,  all  the  quantities  of  which  have  known 
values.  The  result  will  be  the  entropy  of  saturated  steam  for  a 
certain  corresponding  pressure,  to  be  obtained  from  the  steam 
tables.  This  pressure  will  be  that  at  which  the  steam  gives  up  its 
superheat  for  the  example  in  question  and  will  indicate  whether 
the  steam  is  superheated  or  saturated  at  the  end  of  the  flow.  If 
the  pressure  is  greater  than  the  final  pressure  against  which  the 
steam  is  flowing,  it  shows  that  the  steam  becomes  saturated  before 
the  final  pressure  is  reached,  and  hence  will  be  saturated  at  the 
end.  If  the  pressure  determined  is  less  than  the  final  pressure,  the 
steam  will  be  superheated  at  the  end. 

Chart  Showing  Pressures  at  which  Superheated  Steam  Gives 
Up  Its  Superheat. — To  facilitate  calculations  the  chart  in  Fig.  3 
has  been  calculated  by  which  the  pressure  at  which  superheated 
steam  gives  up  its  superheat  in  adiabatic  expansion  can  be  read  off 
directly  and  the  condition  of  the  steam  at  the  end  of  the  expansion 
determined.  Each  curved  line  is  for  a  different  initial  pressure. 
The  pressures  at  the  left  are  those  at  which  steam  superheated  a 
given  amount  gives  up  its  superheat  and  becomes  saturated.  The 
vertical  lines  correspond  to  different  degrees  of  superheat. 

Calculation  of  the  Velocity  of  Flow  of  Superheated  Steam. — 
Having  found  whether  the  given  problem  must  be  solved  by 
equation  (11)  under  Case  I.  or  equation  (12)  under  Case  II. ,  pro- 
ceed as  follows : — 

Case  I. — The  steam  in  this  instance  is  superheated  at  the  end 
of  the  flow.  All  the  quantities  of  equation  (11)  will  be  known  or 
can  be  obtained  from  the  steam  table  except  £S2,  the  final  tempera- 
ture of  the  superheated  steam.  This  must  be  calculated  by  the  aid 
of  characteristic  equation  (16),  Chap.  XII. ,  for  superheated 
steam. 

Case  II. — Here  the  steam  is  saturated  at  the  end  of  its  flow  and 
equation  (12)  will  be  used.  All  the  quantities  of  this  equation  are 
known  except  .r2,  which  is  to  be  obtained  from  characteristic  equa- 
tion (17),  Chap.  XII.,  for  superheated  steam. 

Example  V. — Steam,  superheated  100  degrees  F.,  flows  adiabat- 


170 


100 


150 


140 


130 


\v\ 


\\\\ 


120 


Example;    Steam  of  165  Ib.  abs.  pres.  and  100  superheat 

becomes  saturated  in  expanding  adiabatically 

at  about  66  Ib.  abs. 


lioo 

I 

£90 


70 


50 


40 


10 


100  135 

Degrees  superheat 


150 


175 


200 


Fig    3.      Chart  Showing  Pressures  at  which  Superheated  Steam  Gives  up  its  Superheat 
in    Adiabatic    Expansion.      Specific    Heat    assumedrrO.6. 


276  STEAM  TURBINES 

ically  from  a  pressure  of  135  pounds  absolute  to  a  pressure  of  45 
pounds  absolute.  What  is  its  velocity  of  discharge  ?  Assume  spe- 
cific heat,  c,  to  be  0.6. 

From  the  diagram,  Fig.  3,  we  find  that  steam  superheated  100 
degrees  and  flowing  from  a  pressure  of  135  pounds  to  a  pressure 
of  45  pounds  will  give  up  its  superheat  at  56  degrees  and  hence 
be  saturated  at  the  end  of  the  flow,  making  the  example  come 
under  Case  II.  We  find  x2  from  equation  (17),  Chap.  XII.  ,  as 
follows  :  — 

7*2  /  rx  Ts 


734.  99  r  867.  3 


r  867.3        /  910. 73\     -i 

-  +  . 5027 +  (  2.3  X  0.6  X  log-   -)  -.402 
L810.73        V  810.73/ 


922    L810.73 
=  0.988. 
From  equation  (12)  the  velocity  becomes, 


=  4/50,103(1188.7+60-911.46-243.6) 
=  2166  ft.  per  sec. 

This  velocity  is  only  slightly  greater  than  calculated  for  saturated 
steam  flowing  between  the  same  pressures  in  Example  I.  The 
slight  increase  is  due  to  the  additional  heat  in  the  superheat,  but 
to  partially  offset  this  there  is  more  heat  carried  away  at  the  end, 
in  the  form  of  latent  heat,  since  in  this  example  the  steam  is  more 
nearly  dry  at  the  end  than  in  the  previous  example. 

STEAM  NOZZLE  DESIGN. 

The  fundamental  (and  self-evident)  principle  upon  which  the 
design  of  steam  nozzles  is  based  is  that  the  different  cross-sectional 
areas  of  the  nozzle  must  be  sufficient  to  allow  a  given  weight  of 
the  fluid  to  pass  in  a  given  space  of  time.  In  other  words,  the 
same  weight  of  fluid  must  pass  different  sections  of  the  nozzle  in 
the  same  time. 

This  relation  is  expressed  by  formula  (9)  for  the  weight  of 
flow, 

AV 
W=- 

v 


CALCULATIONS  ON  THE  FLOW  OF  STEAM 


277 


The  Reason  for  Converging  and  Diverging  Nozzles. — Let  us 
examine  this  formula,  and  assume,  first,  that  a  liquid  is  flowing 
through  the  nozzle.  The  specific  volume  v  of  a  liquid  is  constant 
and  hence,  as  the  velocity  V  increases,  owing  to  drop  in  pressure, 
the  area  A  must  decrease ;  or  in  other  words,  the  nozzle  will  con- 
verge, as  in  Fig.  4. 


CONVERGING 
Fig.  4. 


CONVERGING  AND  DIVERGING 
Fig.  5. 


Again,  suppose  steam  to  flow  through  the  nozzle.  In  this  case 
the  specific  volume,  as  well  as  the  velocity,  will  increase  with  the 
drop  in  pressure.  It  is  a  characteristic  of  steam  that  at  first  the 
velocity  increases  more  rapidly  than  the  specific  volume  and  later 
the  specific  volume  increases  more  rapidly  than  the  velocity;  and 
to  conform  to  these  conditions  the  nozzle  should  first  converge  and 
then  diverge,  as  in  Fig.  5. 

Critical  Pressure. — The  point  where  the  ratio 

V 
v 

changes  from  an  increasing  to  a  decreasing  quantity  is  called  the 
critical  point,  and  theoretically  is  at  a  pressure  of  0.58  of  the  ini- 
tial pressure.*  The  pressure  at  the  throat  is  therefore  58  per  cent 
of  the  higher  pressure,  theoretically,  although  tests  in  Chapter  XL 
show  that  it  may  vary  materially  from  this  in  certain  nozzles. 

*Professor  Rateau  of  Paris,  in  Flow  of  Steam  Through  Nozzles,  states  that  he  has 
calculated  the  points  at  which  the  ratio 

V 

v 

becomes  a  maximum  for  a  number  of  different  initial  pressures  and  finds  they  vary 
slightly  with  the  pressure  around  the  value  0.58/>.  It  is  a  very  peculiar  fact  that  this 
relation  exists  so  closely  for  different  pressures.  Mr.  Joseph  C.  Riley,  Massachusetts 
Institute  of  Technology,  Boston,  has  shown  that  the  value  is  affected  somewhat  by 
moisture  in  the  steam.  For  100  pounds  initial  pressure  and  20  per  cent  priming  the 
pressure  at  which  the  ratio  changes  is  0.54 — not  a  wide  variation. 


278  STEAM  TURBINES 

Converging  Nozzle  for  Steam. — In  case  expansion  is  not  car- 
ried far  enough  to  pass  the  critical  point,  where 

V 


begins  to  decrease,  the  diverging  section  of  the  nozzle  is  not  re- 
quired, a  condition  that  exists  when  the  discharge  pressure  is 
58  per  cent  or  more  than  58  per  cent  of  the  initial  pressure,  as  was 
fully  explained  in  Chapter  I.  In  this  case  the  nozzle  of  Fig.  4 
answers  all  requirements. 

Method  of  Procedure  in  Nozzle  Design.  —  Calculations  for 
superheated  steam  are  so  complicated  that  the  method  and  cal- 
culations for  saturated  steam  only  will  be  presented  here.  In  de- 
signing for  superheated  steam  the  same  objects  are  to  be  attained, 
but  we  have  to  use  the  superheated  steam  formulas  for  specific 
volume,  velocity,  etc.,  in  so  far  as  such  formulas  have  been  de- 
veloped and  are  available. 

In  steam  turbine  design  we  know  in  advance  how  many  foot- 
pounds of  energy  per  second  we  wish  to  have  delivered  to  the 
wheel  by  the  steam,  and  what  drop  in  pressure  there  is  to  be  in 
flowing  through  the  nozzle.  Equation  (2)  gives  the  energy  de- 
veloped per  pound  of  steam  for  a  given  drop  of  pressure,  and  the 
total  energy  required  divided  by  the  energy  per  pound  will  give  the 
number  of  pounds  of  steam  per  second,  or  the  fraction  of  a  pound, 
as  the  case  may  be,  that  the  nozzle  must  be  able  to  deliver. 

From  this,  the  area  of  the  nozzle  can  be  determined  by  the  aid 
of  equation  (9).  This  equation  may  be  written 

Wv 

A=—  (13) 

V 

or 


a—  -  (14) 

V 

If  the  nozzle  is  of  circular  cross-section  its  diameter  in  inches  is 
given  by 


576      Wv 

—  X—  (15) 


CALCULATIONS  ON  THE  FLOW  OF  STEAM 


279 


Or,  if  the  designer  prefers,  he  may  use  in  place  of  these  equa- 
tions those  of  Napier  for  the  flow  of  steam,  in  Chapter  XL,  with 
as  good  or  better  results.  The  calculation  of  the  area  of  a  straight 
or  converging  nozzle  is  so  simple  that  no  example  will  be  required. 

Design  of  a  Diverging  Nozzle. — In  a  diverging  nozzle  we  have 
in  addition  to  the  area  of  the  throat  to  find  the  larger  outlet  area. 


140 


l-JO 


NOZZLE 
Fig.   6.     Nozzle   with   Divergence    Carried   too    Far. 

It  has  previously  been  pointed  out,  in  connection  with  the  tests  of 
Chapter  XL,  and  elsewhere,  that  if  the  divergence  of  the  nozzle 
is  carried  too  far,  the  pressure  of  the  steam  will  rise  above  the 
terminal  outside  pressure  before  leaving  the  nozzle  and  there  will 
be  an  attendant  loss  of  energy.  This  is  also  shown  in  Fig  6,  which 
is  a  record  of  an  experiment  by  Francis  Hodgkinson,  and  reported 
in  a  pamphlet  issued  by  the  Westinghouse  Machine  Company. 


280 


STEAM  TURBINES 


The  nozzle  discharges  against  atmospheric  pressure,  but  the  di- 
vergence of  the  nozzle  is  so  great  that  the  steam  reaches  a  pressure 
corresponding  to  that  of  the  atmosphere  shortly  after  the  throat  is 
passed;  and  beyond  this  point  the  steam  expands  to  considerably 
below  atmospheric  pressure,-  reaching  the  lowest  point  at  x.  Then, 
as  it  discharges  to  the  atmosphere,  the  pressure  rises,  and,  of 
course,  the  velocity  decreases.  This  shows  the  importance  of  se- 
curing a  correct  relation  between  the  throat  and  outlet  areas  of  a 
diverging  nozzle.  The  method  to  be  followed  can  best  be  shown 
by  an  example. 

Example  Illustrating  the  Design  of  a  Diverging  Nozzle. — Re- 
quired, the  throat  and  outlet  areas  for  a  nozzle  to  deliver  one 
pound  of  steam  per  second,  flowing  from  a  pressure  of  135  pounds 
absolute  to  a  pressure  of  45  pounds  absolute. 

First,  consider  the  throat  area  on  line  aa,  Fig.  7.  The  steam,  in 
flowing  through  the  converging  inlet  of  the  nozzle,  will  expand  to 
about  %0  of  the  absolute  initial  pressure,  or  to  a  pressure  of  80 


Fig.  7. 

pounds,  approximately.  In  Example  II.,  where  the  steam  ex- 
panded from  135  to  80  pounds,  as  in  the  present  case,  the  velocity 
V  was  found  to  be  1,451.4  feet  per  second.  In  examples  where 
steam  is  known  to  expand  to  %o  °f  the  higher  pressure,  however, 
it  is  not  necessary  to  calculate  the  velocity,  since  for  all  ordinary 
initial  pressures  the  velocity  will  be  practically  constant  and  range 
close  to  1,450  feet  per  second.  (See  Table  I.,  Chapter  XL)  From 
Example  II.,  also,  for  pressure  80  pounds,  ^-=0.966,  while  s,  from 
steam  table=5.43. 


CALCULATIONS  ON  THE  FLOW  OF  STEAM  281 

Hence,  for  section  aa  of  the  nozzle, 

#=80,  F=l,451,  W=l,  and  z/=.™=0.966X  5.43=5.245. 
From  (14), 

144X1X5.245 
a= — 

1,451 

=0.52  square  inch. 

Second,  consider  the  outlet  area  on  line  bb.  Here  the  absolute 
pressure  is  45  pounds  and  in  Example  I.,  where  steam  expanded 
from  135  pounds  to  45  pounds,  as  in  this  case,  we  found  F=2,062 
feet  per  second  and  .r=0.933.  From  the  steam  tables,  ^  for  80 
pounds=9.29. 

Hence,  for  section  bb  of  the  nozzle, 

#=45,  F=2,062,  W—\,  and  z/=*j=0.933X9.?9=8.667. 
144X1X8.667 


a=- 


2,062 
=0.605  square  inch. 

It  will  be  evident  from  the  above  that  the  two  areas  are  directly 
as  the  specific  volumes  and  inversely  as  the  velocities. 

Practical  Considerations. 

As  stated  at  the  outset,  the  calculations  in  this  chapter  are 
based  upon  the  assumption  of  adiabatic  flow.  It  is  a  peculiar 
fact,  however,  that  while  such  calculations  approximate  closely 
to  the  results  of  tests,  it  is  quite  certain  that  the  flow  through 
nozzles  is  not  adiabatic,  or  at  best  it  is  only  approximately  so. 
This  is  indicated  by  the  tests  of  Professor  Lucke  in  Chapter  XL, 
and  is  further  shown  by  the  fact,  which  any  one  can  verify,  that 
the  steam  discharging  from  a  nozzle  does  not  contain  the  amount 
of  moisture  called  for  by  adiabatic  expansion.  It  is  the  experi- 
ence of  experimenters  that  the  discharge  is  either  blue,  indi- 
cating dry  steam,  or  dry  and  white,  indicating  not  over  two  per 
cent  of  moisture. 

Frictional  Losses. — Nozzles  of  different  shapes  and  propor- 
tions have  different  coefficients  of  flow  and  the  best  the  designer 
can  do  in  allowing  for  frictional  losses  is  to  select  his  own  co- 


282  STEAM  TURBINES 

efficients  from  tests  upon  nozzles  as  nearly  as  possible  like  those 
he  contemplates  using.  It  is  apparent  from  the  tests  of  Chapter 
XL  that  a  converging  nozzle  can  be  designed  to  give  a  velocity 
of  discharge  within  two  per  cent  of  the  theoretical  velocity.  The 
diverging  nozzles  show  a  wider  variation,  and  in  some  cases  a 
very  wide  variation,  depending  upon  their  proportions.  In  cal- 
culating the  weight  of  steam  discharged  Napier's  rules  can  be 
used,  as  already  explained,  but  a  proper  coefficient  must  be  selected 
in  each  case  from  tests  upon  similar  nozzles. 

Diverging  Nozzles. — Rosenhain  concludes  from  his  experi- 
ments that  the  taper  of  diverging  nozzles  should  not  be  far  from 
1  in  12  and  that  the  inner  edge  of  the  nozzle  should  be  only 
slightly  rounded.  De  Laval  nozzles  are  made  with  tapers  rang- 
ing from  about  1  in  10  to  1  in  20,  with  inlet  only  slightly 
rounded.  The  experiments  of  Mr.  Strickland  L.  Kneass,  engineer 
of  the  injector  department  of  William  Sellers  &  Co.,  Philadelphia, 
indicate  that  better  results  are  obtained  from  nozzles  with  well- 
rounded  inlets  and  a  taper  of  about  1  in  6 ;  and  that  if  properly 
proportioned  it  is  possible  to  secure  a  velocity  of  flow  within  two 
per  cent  of  theoretical.  One  of  several  series  of  experiments, 
records  of  which  have  been  furnished  the  author  by  Mr.  Kneass, 
was  upon  five  different  nozzles  having  a  divergent  taper  of 
1  in  6,  but  with  a  ratio  of  discharge  to  throat  areas  carefully  cal- 
culated for  different  initial  pressures.  The  nozzles  discharged  at 
atmospheric  pressure  against  a  parabolic  target,  which  deflected 
the  steam  through  an  angle  of  90  degrees.  The  target  was  con- 
nected with  a  delicate  weighing  device  by  which  the  impact  of 
the  steam  could  be  accurately  determined. 

Of  the  several  nozzles,  one  designed  for  30  pounds  initial 
pressure,  gauge,  and  having  a  throat  diameter  of  4.14  mm  and 
discharge  diameter  of  4.32  mm,  gave  the  best  average  results 
throughout  the  whole  range  of  pressures,  as  per  table  below : — 

Impact  Pressures  Weight  Discharged, 

Initial  Pressure,  in  Ib.  per  sq.  Ib.  per  sq.  mm 

Gauge.  mm  of  Nozzle.  of  Nozzle. 

120  0.2401  0.002870 

90  0.1783  0.002249 

60  0.1174  0.001648 

30  0.0591  0.000965 

15  0.0299  0.000688 


CALCULATIONS  ON  THE  FLOW  OF  STEAM  283 

From  these  results  the  author  has  calculated  the  actual  veloci- 
ties of  discharge,  using  the  formula  employed  in  Rosenhain's 
tests,  and  compared  them  with  the  theoretical  velocities,  taken 
from  Chart  No.  3  in  the  appendix,  as  follows : — 

Gauge 
Pressure. 

120 
90 
60 
30 
15 

These  results  show  that  at  30  pounds,  for  which  the  nozzle  was 
designed,  the  loss  is  only  1.5  per  cent,  and  by  comparing  results 
for  the  whole  range  of  pressures  it  will  be  evident  that  a  nozzle 
is  more  efficient  when  designed  for  too  low  a  pressure  than  for 
too  high  a  pressure. 

From  the  velocities  given  above  the  actual  and  theoretical 
energy  of  the  steam  jet  may  also  be  calculated  and  compared. 
The  energy  loss  in  the  above  cases  will  be  found  to  range  from 
3  to  13  per  cent. 


Actual 

Theoretical 

Loss, 

Velocity. 

Velocity. 

Difference. 

Per  Cent. 

2,690 

2,820 

130 

.046 

2,550 

2,650 

100 

.038 

2,290 

2,400 

110 

.045 

1,970 

2,000 

30 

.015 

1,400 

1,500 

100 

.066 

CHAPTER  XIV 

TURBINE  VANES. 

The  Vanes  of  Impulse  Turbines. 

In  preceding  chapters  we  have  studied  the  principles  of  the  flow 
of  steam  and  the  conversion  of  the  heat  energy  of  steam  into  the 
mechanical  kinetic  energy  of  the  escaping  jet.  The  subject  now  to 
be  considered  is  the  transference  of  this  mechanical  energy  of  the 
jet  to  the  vanes  of  the  turbine  wheel. 

At  the  beginning  of  Chapter  I.  the  meaning  of  absolute  and 
relative  motion  was  explained,  as  related  to  the  action  of  a  fluid 
upon  a  moving  vane,  and  the  treatment  there  given  will  furnish 
sufficient  introduction  for  what  is  to  follow. 

Diagram  for  a  Moving  Vane. — In  Fig.  1  herewith  a  turbine 
vane  moves  in  the  direction  of  the  horizontal  arrow  and  is  acted 
upon  by  a  jet  of  steam  flowing  in  the  direction  of  the  inclined 


Fig.  1. 

arrow.  The  parallelograms  of  motion  show  the  velocity  and 
direction  of  motion  of  the  steam  in  entering  upon  and  leaving  the 
vane ;  and  the  velocity  and  direction  of  motion  of  the  vane  itself. 
The  lengths  of  the  several  lines  represent  velocities  in  feet  per 
second,  drawn  to  any  convenient  scale.  In  the  diagram, 

V  is  the  initial  and  v  the  final  absolute  velocity  and  direction  of 
the  steam ; 


TURBINE  FANES  285 

R  is  the  initial  and  r  the  final  velocity  of  the  steam  relative  to 
the  vane. 

w  is  the  velocity  and  direction  of  motion  of  the  vane. 

If  there  were  no  loss  through  friction  or  eddying,  the  relative 
velocities  R  and  r  would  be  equal,  and  this  will  be  assumed  the 
case  in  what  follows,  unless  otherwise  stated. 

A  is  the  angle  made  by  V  with  the  direction  of  motion  of  the 
vane;  C,  the  angle  made  by  R;  and  D  and  B,  the  corresponding 
angles  made  by  the  steam  when  leaving  the  vane.  In  what  fol- 
lows A  and  D  will  be  designated  as  the  "initial"  and  "final"  angles, 
and  C  and  B  the  "entrance"  and  "exit"  angles. 

For  tangential  action  upon  the  vane,  allowing  the  steam  to 
enter  upon  it  without  impact  and  to  leave  it  without  commotion, 
the  vane  should  be  tangent  to  R  at  the  entrance  and  tangent  to  r 
at  the  exit.  The  shape  of  the  vane  between  the  entrance  and  exit 
is  not  very  important,  so  long  as  the  curve  is  gradual  and  smooth. 

In  hydraulic  turbines  the  water  usually  flows  either  outward  or 
inward,  in  a  direction  generally  radial ;  and  as  the  vanes  are  large 
it  is  necessary  to  take  into  account  the  difference  in  velocity  of 
their  inner  and  outer  circumferences  when  proportioning  the 
angles,  etc.  In  steam  turbine  work  this  is  not  necessary,  since  the 
steam  usually  flows  in  an  axial  direction  and  it  is  sufficiently 
accurate  to  assume  that  the  vane  moves  forward  in  a  straight  line 
at  a  speed  equal  to  that  of  the  mean  circumferential  speed  of  the 
vanes. 


Fig.  2. 


Calculating  the  Parts  of  the  Diagram. — Fig.  2  shows  a  con- 
venient arrangement  of  the  velocity  diagram  for  finding  the  values 
of  the  different  elements,  either  by  graphical  construction  or  by 


286  STEAM  TURBINES 

calculation.  The  several  lines  are  lettered  to  correspond  with 
Fig.  1  and  the  several  parts  can  be  calculated  by  the  simple 
formulas  of  trigonometry.  The  most  important  formula  used  is 
the  one  stating  that  "In  any  triangle  the  square  of  any  side  is 
equal  to  the  sum  of  the  squares  of  the  other  two  sides,  minus 
twice  their  product  into  the  cosine  of  their  included  angle." 

For  example,  if  there  are  given  values  of  V  ,  w,  and  angles  A 
and  B;  and  it  is  required  to  find  R  (  =r),  v  and  angles  C  and  D, 

then  R*=w*+y*—  2  w  V  cos  A  •    (1) 

Z,2_.r2_j_w2_2  r  w  cos  B  (2) 

Also,  it  can  be  shown  that 

V  cos  A  —  w 

Cos  C=  -  —  (3) 

K 

r  cos  B  —  w 

Cos  D=  —  -  (when  line  v  is  inclined  to  the  right  as  in 

v 

Fig.  2).  (4) 

w  —  r  cos  B 

Cos  D=  -         -  (when  line  v  is  inclined  to  the  left).      (5) 
v 

Turbine  Efficiency.  —  The  energy  of  one  pound  of  steam  imping- 
ing against  the  turbine  vanes  is 

V2 


and  of  one  pound  of  steam  as  it  leaves  the  vanes  is 


The  energy  absorbed  by  the  vane  is  therefore 

F2- 


By  the  principle  of  machines, 

energy  absorbed  by  vanes 


Efficiency^ 


F2 
that  velocities  only  need  be  considered. 


total  energy  delivered  to  vanes 

F2— z/2 

,  the  2g  canceling  in  each  case  so 


TURBINE  VANES  287 

But  by  trigonometry, 

y*=w2+R2—2  w  R  cos  (180— C), 

which  reduces  to 

F2— w2+7?2+2  w  R  cos  C. 
Also,  v2=w2+r2— 2  w  r  cos  B, 

whence, 

~  cos  C — (w2-\-r2 — 2  iv  r  cos  5) 

-  (i} 

(8) 


F2  F2 

R2 — r2+2  w  (.ft  cos  C+r  cos 

V2 

Assuming  that  R~r,  as  previously  explained, 
*—v*     2  w  R  ( cos  C+cos  B  ) 


Efficiency=- 


V 


Conditions  of  High  Efficiency. — It  is  obvious  that  to  attain  high 
efficiency  the  final  absolute  velocity  v  of  the  steam  must  be  small, 
as  otherwise  energy  would  be  wasted  without  doing  work  on  the 


Fig.  3. 

vanes.  In  Fig.  3  v  has  a  small  value  and  it  will  be  evident  that 
this  is  partly  due  to  the  fact  that  angle  B  is  small  and  partly  that 
w  and  r  are  equal.  A  few  minutes  spent  in  drawing  diagrams 
will  demonstrate  that  if  angle  B  is  large,  or  if  w  and  r  differ  con- 
siderably in  length,  the  value  v  will  be  materially  increased. 

In  laying  out  a  diagram  economical  conditions  may  be  attained 
by  selecting  angles  A  and  B  as  small  as  practicable,  say  from  15 
to  20  degrees, .and  angle  C=2A,  thus  making  w  and  R  equal. 


288  STEAM  TURBINES 

The  value  of  w,  and  hence  of  R,  will  then  be  calculated  by 

V 

w=—  (10) 

2  cos  ,4 

The  final  absolute  velocity  is  calculated  by 


(1— cosB)  (11) 

While  these  proportions  do  not  give  quite  the  maximum  effi- 
ciency, they  produce  a  simple  construction  and  are  satisfactory  if  it 
is  possible  to  use  them.  Steam-  velocities  are  so  high,  however, 
that  the  wheel  velocity  w  must  usually  be  selected  from  considera- 
tions of  safety  and  utility  rather  than  of  theory. 

The  question  of  efficiency  can  be  made  clear,  without  difficult 
calculations,  by  reference  to  the  graphical  constructions  of  Figs. 
3,  4,  and  5. 

Efficiency  as  Shown  by  Diagrams. — 

Example  /.—Let  A  and  B  =20  degrees  and  F=3,000  feet  per 
second. 

In  accordance  with  the  last  article,  let  C=2^4=40  degrees ; 

V          3,000 

w= = =1,600   feet  per   second.     This   is  also   the 

2  cos  A     2X-94 

value  of  R. 

In  Fig.  3  draw  w  to  any  suitable  scale,  to  represent  1,600,  and 
V  at  an  angle  of  20  degrees  with  w  to  represent  3,000.  Then  draw 
R=w  at  an  angle  of  40  degrees  with  the  latter.  Complete  the 
parallelogram. 

Now  construct  the  parallelogram  for  the  discharge  in  the  same 
manner,  making  r=zv  and  angle  5=20  degrees. 

v  is  the  absolute  velocity  of  discharge  and  the  vane  curve  is 
drawn  with  the  entrance  and  exit  surfaces  tangent  respectively  to 
R  and  r. 

It  will  be  evident  that  a  wheel  with  vanes  laid  out  as  in  Fig.  3 
will  have  a  high  efficiency. 

Example  2. — Fig.  4  has  been  constructed  to  show  the  influence 
of  the  initial  angle  A  upon  the  value  of  v.  As  in  the  previous 
diagram  C=2A,  w—R  and  5=20  degrees.  But  angle  A  has  been 
made  45  instead  of  20  degrees  and  C  90  instead  of  40  degrees. 


TURBINE  VANES 


289 


The  result  is  that  w  and  R  and  hence  r  are  much  greater  than 
before  and  v  in  consequence  is  greater  and  the  wheel  will  be  less 
efficient.  It  will  be  noted,  however,  that  while  the  size  of  angle  A 
in  Fig.  4  is  more  than  double  its  size  in  Fig.  3,  the  value  of  v  is 


Fig.  4. 

increased  only  a  few  per  cent.  There  can  thus  be  a  considerable 
latitude  in  the  selection  of  the  initial  angle.  The  chief  disadvan- 
tage of  a  large  angle  in  steam  turbine  work  is  that  it  necessitates 
a  high  value  for  the  speed  w  of  the  wheel  circumference ;  and  if 
we  attempt  to  reduce  the  wheel  velocity  and  still  maintain  a  large 
initial  angle,  the  result  is  not  good,  as  the  next  example  will 
show. 

Example  j. — In  Fig.  5  the  angle  A  was  made  45  degrees  as 
in  the  last  example,  and  the  angle  B  20  degrees,  as  in  both  Exam- 
ples 1  and  2.  Instead  of  selecting  C=2A}  however,  the  speed  w 
of  the  wheel  was  kept  the  same  as  in  Example  1,  or  less  than  in 


Fig.   5. 


Example  2,  and  the  angle  C  was  then  determined  by  drawing  the 
parallelogram.  The  final  result  is  a  value  for  v  greater  than  in 
either  of  the  previous  cases,  indicating  that  the  construction  is  not. 
so  good  as  where  C=2A  and  w=R,  as  in  Figs.  3  and  4. 


290 


STEAM  TURBINES 


While  the  foregoing  examples  are  not  of  the  nature  of  demon- 
strations, they  indicate  why  it  is  desirable  to  ru'.ve  angle  B  as 
small  as  possible  and  A  reasonably  small  ;  while  the  sides  of  both 
parallelograms  should  be  equal,  making  C=2A 

Vanes  with  Entrance  and  Exit  Angles  Equal.  —  An  important 
case  for  impulse  steam  turbines  is  that  of  symmetrical  vanes  hav- 
ing entrance  and  exit  angles  equal.  With  vanes  so  proportioned 
there  is  no  thrust  to  be  taken  care  of,  due  to  the  reaction  of  the 
steam  leaving  the  vanes,  since  the  reaction  is  balanced  by  the  im- 
pulse of  the  jet  striking  the  vanes. 


Fig.  6. 

In  Fig.  6  is  the  diagram  for  this  construction.  Let  us  assume 
F=4,000  feet  per  second,  A— 20  degrees,  C=2^4=40  degrees 
(the  condition  of  high  efficiency),  and  BJ  which  is  equal  to  C  to 
make  the  vane  symmetrical,  is  also  40  degrees. 

By  formula  (10), 

4,000 

w=—     —=2,128   feet  per  second.     This  is  also  the  value  of 
2X-94 

R  and  r. 

By  formula  (11), 

77=2,128 v'2X- 234=1,456  feet  per  second. 

The  efficiency  may  be  calculated  by  e.'ther  the  first  or  second 
parts  of  formula  (9)  as  most  convenient.  Taking  the  first  part, 
we  have, 

(4,000)2— (1,456)2 

Efficiency^—  — =86.7  per  cent. 

(4,000)2 

Obviously  no  such  wheel  velocity  as  the  above  would  be  possi- 


TURBINE  VANES  291 

ble  and  no  velocity  of  steam  as  great  as  4,000  feet  per  second 
would  usually  be  attained,  although  it  is  sometimes  reached  in  the 
De  Laval  turbine.  Impulse  turbines  are  ordinarily  of  the  multi- 
cellular  type  and  the  drop  in  pressure  between  any  two  compart- 
ments would  not  be  more  than  enough  to  produce  the  maximum 
velocity  of  flow  from  a  converging  nozzle,  or  about  1,450  feet 
a  second.  If  there  were  one  wheel  in  each  compartment  the 
velocity  of  its  vanes,  calculated  on  the  same  basis  as  the  last 
example,  would  be  771  feet  a  second,  which  is  within  feasible 
limits. 

Formulas  for  Highest  Efficiency.  —  It  has  been  explained  that 
while  the  conditions  just  outlined  produce  a  wheel  of  high  effi- 
ciency, they  do  not  produce  the  highest  efficiency  possible.  In 
notes  issued  to  his  students,  Prof.  Edward  F.  Miller,  Massachu- 
setts Institute  Technology,  has  worked  out  formulas  for  the 
theoretical  highest  efficiency  for  symmetrical  vanes.  The  demon- 
strations are  somewhat  difficult  and  only  the  results  will  be  given. 
It  is  assumed  that  A  and  V  are  known.  Then, 

=2tan^  (12) 

V  cos  A 
w=—  (13) 


smA  (14) 

Efficiency=cos  A  (15) 

The  calculated  results  for  the  example  of  the  last  article  are 
C=36  degrees,  plus;  w  =1,879  feet  per  second;  z/=l,368  feet  per 
second;  efficiency=88.3  per  cent.  These  may  be  compared  with 
the  results  found  by  the  other  method. 

Efficiency  of  Pelton  Type  Wheel.  —  In  Fig.  7  is  a  Pelton  bucket 
moving  in  the  direction  K  with  a  velocity  w.  The  jet  strikes  the 
bucket  in  the  direction  in  which  it  is  moving  as  indicated  by 
arrow  A. 

If  the  bucket  were  so  shaped  that  there  were  a  complete  re- 
versal of  the  stream,  allowing  it  to  escape  in  the  opposite  direction 
from  which  it  entered  the  bucket,  as  indicated  by  the  dotted  line, 
the  efficiency  would,  neglecting  friction,  depend  only  on  the 


292  STEAM  TURBINES 

velocity  of  the  wheel.  If  the  wheel  traveled  with  half  the  velocity 
of  the  jet  the  efficiency  would  be  100  per  cent,  and  as  the  velocity 
of  the  wheel  decreased  the  efficiency  would  grow  less. 


Fig.  7. 

This  is  an  impossible  condition,  of  course,  and  the  buckets  must 
be  designed  to  allow  the  fluid  to  depart  at  a  slight  angle  from  the 
path  of  the  bucket,  as  indicated  by  r  in  Fig.  7. 

Let  Wj  Fig.  7,  be  the  wheel  velocity.  If  V  is  the  velocity  of  the 
jet,  then  r,  the  relative  velocity  of  the  fluid  on  the  bucket,  is  V — w. 
The  final  absolute  velocity  v  is  calculated  by  formula  (2), 

v2=r2-^ w2— 2  r  w  cos  B, 
where  B  is  the  angle  of  discharge. 

Having  found  this  the  efficiency  is  easily  calculated  by  the.  first 
part  of  the  formula  (9)  ;  or  the  second  part  of  the  formula  may  be 
used  remembering  that  angle  C  is  zero  and  that  cos  0=1. 

Example. — Let   F=l,000  feet  per  second;  w=400  feet;  and 
5=15  degrees.    Then  r=F — w=600  feet  per  second  and 
*;2:=360,000+1GO,000— 2X600X400X.966 
=56,320. 

1,000,000—56,320 

Efficiency^ —  —=94.4  per  cent. 

1,000,000 

This,  like  other  calculations  for  efficiency  in  this  chapter,  is. 
purely  theoretical  and  is  higher  than  can  be  realized  in  actual  con- 
ditions. 

Diagrams  for  Compound  Impulse  Turbines. — In  Fig.  8  the  ini- 
tial velocity  V^  of  the  steam  is  assumed  to  be  four  times  the  wheel 
velocity  w,  and  angle  A1=20  degrees.  The  steam,  having  acquired 
its  velocity,  flows  of  its  own  momentum  through  two  sets  of  mov- 


TURBINE  VANES 


293 


ing  vanes  and  one  set  of  guide  vanes  between  them,  and  finally 
issues  at  an  absolute  velocity  v.2.  To  avoid  end  thrust  the  moving 
vanes  must  be  symmetrical  and  hence  C1=5]  and  C2=B2.  If  there 


Fig.  8. 


Fig.  9. 


were  no  loss  through  friction,  etc.,  then  the  entrance  velocity  Rl 
relative  to  the  vane  would  equal  relative  velocity  ^  at  exit;  the 


294  STEAM  TURBINES 

absolute  velocity  v^  of  steam  leaving  the  first  set  of  moving  vanes 
would  equal  the  absolute  velocity  V2  in  passing  through  the  inter- 
mediate set  of  guide  vanes ;  and  relative  velocities  R2  and  r2  would 
be  equal.  The  dotted  arcs  indicate  which  velocities  are  to  be 
drawn  equal  in  this  construction. 

Case  Where  There  is  Loss  Through  Friction. — If  it  is  desired 
to  take  account  of  frictional  loss  this  may  be  done  as  in  Fig.  9, 
where  relative  velocity  1\  is  made  less  than  relative  velocity  R^; 
absolute  velocity  V2  less  than  v± ;  and  relative  velocity  r2  less  than 
R2.  The  diminution  of  the  velocities  may  be  made  either  an 
arbitrary  amount  in  each  case  or  a  certain  percentage  of  the 
velocity,  as  desired. 

It  will  be  noted  that  in  Figs.  8  and  9  the  vane  angles  are  differ- 
ent for  the  two  wheels.  If  desired  to  make  them  the  same  for 
convenience  in  manufacture,  a  plan  must  be  followed  similar  to 
that  now  to  be  described  in  connection  with  reaction  blades. 

The  Vanes  of  Reaction  Turbines. 

In  Fig.  10  is  a  diagram  by  which  the  action  of  the  steam  in 
reaction  turbines  may  be  studied.  For  convenience  in  manu- 
facture the  guide  and  moving  vanes  in  any  one  step  or  series  of 
the  turbine  are  usually  made  alike.  The  upper  parallelogram 
shows  the  absolute  velocity  and  direction  V  of  steam  leaving  the 
guide  vanes,  its  velocity  and  direction  R  relative  to  the  moving 
vanes  at  the  point  of  entrance,  and  the  velocity  and  direction  w  of 
the  moving  vanes. 

The  lower  parallelogram  shows  the  absolute  velocity  and  direc- 
tion v  of  the  steam  leaving  the  moving  vanes,  its  velocity  and 
direction  r  relative  to  the  moving  vanes  at  the  point  of  exit,  and 
the  velocity  and  direction  w  of  the  moving  vanes. 

Characteristics  of  the  Reaction  Diagram. — The  essential  dif- 
ference between  this  diagram  and  those  for  impulse  turbines  is 
that  the  relative  velocity  r  is  greater  than  R.  In  the  reaction  tur- 
bine steam  first  expands  and  acquires  a  velocity  in  the  guide  pas- 
sages, as  in  the  impulse  turbine.  Then,  in  flowing  through  the 
wheel  passages  it  continues  to  expand  and  acquires  a  greater 
velocity  relative  to  the  moving  vanes  than  it  had  at  the  entrance. 
The  absolute  velocity  of  the  steam,  of  course,  diminishes. 


TURBINE  VANES  295 

Call  M  the  velocity  acquired  by  the  steam  as  a  result  of  the 
expansion  in  the  passages  of  any  one  row  of  vanes.  Then,  since 
steam  enters  each  set  of  guide  passages  with  an  initial  velocity 
equal  to  the  velocity  v  of  the  steam  leaving  the  wheel  vanes,  we 
have :  Velocity  V  of  steam  leaving  the  guide  vanes=z/+w. 

In  the  moving  vanes  the  velocity  «  acquired  through  expansion 
has  the  effect  of  increasing  the  velocity  of  the  steam  relative  to 
the  vanes,  so  that  relative  velocity  r  of  steam  leaving  the  moving 
vanes=^+w.  From  the  foregoing  we  therefore  have,  V— v— 
r — R,  or  the  difference  between  the  absolute  velocities  V  and  v 
equals  the  difference  between  the  relative  velocities  r  and  R. 


Fig.  10. 

Construction  of  a  Reaction  Diagram. — In  Fig.  10,  where  the 
guide  and  moving  vanes  are  alike,  it  will  be  evident  that  angles 
A  and  B  must  be  equal  and  angles  C  arid  D  equal,  to  secure  tan- 
gential action  upon  the  vanes.  Hence,  in  Fig  11  we  make  these 
angles  equal,  having  A  and  B  as  small  as  practicable  and  making 
angles  C  and  D  from  70  to  90  degrees.  In  hydraulic  work  90  de- 
grees is  frequently  selected  for  this  angle. 

Referring  to  triangles  (1)  and  (2),  Fig.  11,  we  have  the  com- 
mon side  w  and  angles  A  and  B  equal ;  consequently  these  trian- 
gles are  equal  and  V—r  and  R=v.  Having  assumed  a  suitable 
value  for  w,  and  knowing  the  angles,  the  other  sides  may  be  cal- 
culated by  the  principle  in  trigonometry  that  in  any  triangle  the 


296 


STEAM  TURBINES 


Fig.  11. 

sides  are  proportional   to  the  sines  of  the  opposite  angles,  as 
follows : — 


R=- 


w  sin  (180 — C)        w  sin  C 
sin  (C—A)     =  sm(C—A) 
w  sin  A 


sin  (C—A) 

iv  sin  B 
sin  (D—B) 

«»  sin  D 
'sin  (D—B) 


(16) 
(17) 
(18) 
(19) 


The  efficiency  of  the  reaction  turbine  must  be  studied  by  taking 
the  machine  as  a  whole,  since  the  action  is  continuous  from  the 
first  to  the  last  row  of  vanes  and  the  losses  through  leakage,  fric- 
tion, etc.,  are  such  that  no  estimate  of  efficiency  can  be  made  by 
calculating  the  efficiency  of  any  one  set  of  guide  and  moving  vanes. 
The  efficiency  can  only  be  determined  by  experiment. 


TURBINE  VANES 


TESTS  UPON  BUCKETS  AND  CHANNELS. 


297 


Tests  of  Strickland  L.  Kneass,  C.  E. — In  1894  a  small,  experi- 
mental turbine  was  built  and  tested  in  the  injector  department  of 
William  Sellers  &  Co.,  Inc.,  Philadelphia.  The  machine  was  so 
designed  that  it  could  be  assembled  to  operate  on  the  plans  of 
several  types  of  turbines,  including  the  Parsons,  what  later  became 
known  as  the  Curtis,  and  the  De  Laval ;  and  when  under  the  latter 
arrangement  it  could  be  run  either  as  a  compound  or  a  simple  tur- 
bine. 

Preliminary  to  the  turbine  tests  an  investigation  was  undertaken 
of  the  action  of  steam  jets  upon  vanes  and  in  flowing  through 


Fig.   12.     Flat  and   Parabolic   Targets   for   Measuring   Impulse. 

curved  tubes.  These,  as  well  as  the  turbine  tests,  were  at  the  in- 
ception of  Mr.  Kneass,  whose  experimental  work  has  several 
times  before  been  referred  to,  and  who  has  allowed  the  author  to 
make  selections  from  the  records  of  tests  on  buckets  and  tubes. 

In  the  preliminary  tests  a  delicately  balanced  target  was  em- 
ployed, against  which  a  steam  jet  from  a  nozzle  was  directed.  The 
target  was  provided  with  a  sensitive  weighing  device  and  was  so 
manipulated  in  taking  readings  as  to  eliminate  the  effect  of  fric- 
tion in  the  final  results.  In  order  to  measure  the  impulse  of  the 
jet,  the  target  was  made  with  a  parabolic  surface  coming  to  a 
point  at  the  center,  so  as  to  deflect  the  stream  through  an  angle  of 
ninety  degrees,  with  as  little  loss  as  possible.  When  vanes  or 
passages  of  any  particular  shape  were  to  be  tested,  the  piece  com- 


298 


STEAM  TURBINES 


prising  these  was  bolted  to  the  target,  by  which  means  the  im- 
pulse or  reaction  could  be  measured. 

Impact  vs.  Tangential  Action. — In  theoretical  discussions  of 
turbine  vanes,  it  is  assumed  that  the  fluid  must  glide  upon  them 
tangentially  in  order  to  avoid  losses  from  impact.  In  operation, 
however,  hydraulic  turbines  seldom  run  at  exactly  the  speed  re- 
quired to  produce  tangential  action ;  and  in  fact,  tests  have  not  in- 
frequently shown  that  the  best  results  are  obtained  by  running 
at  slightly  faster  or  slower  speeds. 

In  elucidation  of  this  subject  are  tests  by  Mr.  Kneass,  in  which 
jets  were  allowed  to  discharge  against  flat  and  parabolic  targets, 
as  in  Fig.  12,  and  the  pressure  measured  in  each  case,  with  results 
as  in  Table  I. 

TABLE  I. 


FLAT  TARGET. 

PARABOLIC  TARGET. 

Steam 
Pressure, 
Gauge. 

Distance 
A. 

Pressure  on 
Target  per  sq. 
Millimeter  of 
Nozzle,  Pounds 

Steam 
Pressure, 
Gauge. 

Distance 
A. 

Pressure  on 
Target  per  sq. 
Millimeter  of 
Nozzle,  Pounds 

120 
120 
120 
120 
120 

Average, 

|  inch 
J  inch 
i  inch 
1    inch 
1£  inch 

0.2302 
0.2:ttU 
0.2390 
0.2-161 
0.2461 

120 
90 
60 
30 

|  inch 
\  inch 
iinch 
|  inch 

0.2413 
0.17^7 
0.1109 
0.0540 

0.2387 

These  tests  show,  that  by  adjusting  the  distance  of  the  nozzle 
from  the  target  it  was  possible  to  secure  as  great  a  pressure  with 
the  flat  as  with  the  parabolic  target  and  point  to  the  conclusion 
that  perfect  tangential  action  of  a  jet  upon  a  vane  is  not  essential 
to  high  economy. 

Another  test  illustrating  the  same  fact  was  made  with  two  dif- 
ferent vanes,  Fig.  13,  one  with  an  entrance  angle  of  60  de- 
grees and  one  with  an  entrance  angle  of  20  degrees.  The 
angle  of  the  nozzle  and  the  exit  angle  of  the  vanes  were  the  same 
in  both  tests.  The  steam  pressure  was  15  pounds,  gauge,  an<5 
the  pressures  upon  the  target  were,  for  the  60-degree  angle 


TURBINE  VANES 


299 


0.0454  pound  and  for  the  20-degree  angle  0.0475  pound,  the  jet 
flowing  on  tangentially  at  the  latter  angle.  When  the  entrance 
angle  was  changed  from  20  degrees  to  60  degrees,  there  was  a 
falling  off  of  only  4.5  per  cent. 


Fig.  13.     Experiment   with    Different    Entrance    Angles. 


Experimental  Buckets. — Among  the  other  buckets  tested  were 
those  of  Fig.  14  and  selected  results  of  tests  upon  them  are  given 
in  Table  II.,  which  requires  no  explanation,  since  the  reader  can 
easily  make  his  own  comparison  of  data. 

Owing  to  the  falling  off  in  the  pressure  on  the  target  with 
bucket  No.  6,  as  compared  with  bucket  No.  1,  another  experiment 
was  made  with  a  bucket  of  similar  design.  The  nozzle  used  with 
No.  6  was  only  4  mm.  diameter,  while  that  used  with  No.  1  was 
6  mm.  diameter.  In  the  additional  experiment  the  nozzle  was  7 
mm.  diameter,  and  beveled  off  as  in  the  No.  6  test,  but  with  a 
little  larger  passage  leading  to  the  nozzle  proper.  This  gave  a 
pressure  against  the  target  of  0.0572  pound  per  square  millimeter 
of  nozzle  area,  with  an  initial  pressure  of  15  pounds,  gauge, 
showing  a  gain  of  17  per  cent  over  the  experiment  with  the  4  mm. 
nozzle. 


300 


STEAM  TURBINES 


Experiments  with  Curved  Tubes. — A  series  of  experiments  was 
made  with  curved  tubing  arranged  as  at  A  and  B  in  Fig.  15 


No.  1 


Fig.  14.     Experimental    Buckets. 

which  shows  what  large  losses  may  result  from  friction.  Copper 
tubing  was  used,  y%  inch  diameter,  bent  on  an  inner  radius  of  Y$ 
inch.  The  nozzle  used  was  1  in  6  taper,  4.14  mm.  diameter  at  the 
throat  and  7.5  mm.  diameter  at  the  mouth.  Tests  were  first  made 
with  a  single  tube  mounted  on  the  target  and  bent  through  an 
angle  of  180  degrees.  Three  similar  bends  were  next  used  with  the 
first  two  stationary  and  their  ends  separated  %6  inch,  as  at  B, 
and  the  third  one  mounted  on  the  target.  The  results  with  this  ar- 
rangement, when  compared  with  the  first  set  of  results,  show  the 
losses  in  pressure  due  to  friction  in  the  tubes  and  disturbance 
caused  by  the  spaces  between  the  ends  of  the  tubes.  Finally,  tests 
were  made  with  a  single  tube  mounted  on  the  target,  (this  tube 
not  shown)  making  lJ/£  turns,  or  the  same  number  of  turns  all 


TURBINE  VANES 
TABLE  II. 


301 


og 

"»   £ 

M 

a? 

0  *  U     . 

jj 

.0  U 

*«fi 

H 

fl§d 

S  ft,  8  c 

Remarks. 

§  a 

5  fl 

Cfl  HH 

88  a 

gp!* 

SW 

5.S 

OnhJi^ 

No.  1 

6 
6 
6 
6 
6 

i 

CltL 

60 
120 
15 

.0484 
.0996 
.1940 
.3659 
.0607 

Received  steam  nicely  into  bucket,  no  back- 
ward discharge.     At  120  pounds'  pressure 
steam  escaped  in  all  directions  from  bucket. 
At  other  pressures  there  was  little  discharge 
from  sides,  and  steam  left  in  a  good  jet. 

No    2 

G 

A 

^ 

.04^0 

All  steam  entering  channel. 

No.  3 

0 

6 

s 

Kl 

.0480 
.0928 

Apparently  no  resistance  in  striking  knife 

6 

s 

CO 

.1930 

edge. 

6 

0 

15 

.0538 

Bucket  had  open  ends.    Steam  left  bucket  in 

No.  4 

6 

& 

15 

.0530 

a  nice  double  jet.    Knife  edge  was  cut  in 

0 

il 

15 

.0500 

the  short  time  of  the  experiments. 

G 

.0572 

Same  as  No.  4,  but  ends  of  bucket  were  en- 

6 

A 

.0620 

closed,  thus  confining  the  steam  sideways. 

No.  6 

4 
4 

Close 
Close 

15 
15 

.0475 
.0419 

Discharged  all  into  B. 
Discharged  equally  into  A  and  B. 

told  as  the  three  tubes  at  B  in  the  diagram.  The  pressures  upon 
the  target  in  the  three  cases,  per  square  millimeter  of  nozzle  area, 
are  given  in  Table  III. 


A  Nozzle 


Nozzle 


Target  ^        Target 

Fig.   15.     Experiments    with    Curved    Tubing. 

Experiments  ^v^th  Rectangular  Tubes. — Instead  of  continuing 
experiments  with  the  round  tubing,  a  curved  rectangular  tube  was 
made  with  three  half  turns,  as  at  B,  Fig.  16.  The  dimensions  of 


302 


STEAM  TURBINES 

TABLE  III. 

PRESSURES  IN  POUNDS  PER  SQUARE  MILLIMETER  OF  NOZZLE,  WITH 
TUBING,  ARRANGED  AS  IN  FIG.  15. 


Steam  Pressure, 
Gauge. 

Single  Half 
Turn. 

Three  Half  Turns 
—  Pressure  on 
Last  One. 

Continuous  Coil, 
1£  Turns. 

12) 

0.400 

0.248 

0.371 

90 

0.2>8 

0.170 

0.260 

60 

0.174 

0.098 

0.159 

30 

0.084 

0.04:3 

0.066 

the  tube  increased  from  8.2  by  8.2  millimeters  at  the  inlet  to  8.75 
by  12  millimeters  at  the  outlet. 

The  nozzle  used  in  this  series  was  of  1  in  6  taper,  4.14  mm. 
diameter  at  the  throat  and  G.5  mm.  at  the  mouth.  It  is  the  same 
nozzle  used  in  the  tests  in  Table  I.  with  the  parabolic  target.  Four 
arrangements  were  tried  as  illustrated  at  ^/,J9,  C,and  D  in  Fig.  16 


Fig.  16.     Experiments  with  Rectangular  Curved  Tubing. 


TURBINE  VANES 


303 


First,  tests  were  made  with  a  continuous  bend,  B,  mounted  on  the 
target,  with  results  tabulated  in  column  3,  Table  IV.  Second,  the 
three  half-bends  were  cut  apart,  as  at  C,  with  openings  %4  inch 
wide,  and  all  three  bends  located  on  the  target,  the  results  being 
tabulated  in  column  4  of  the  table.  Third,  a  single  half-turn, 
shown  at  A,  was  used,  the  results  being  given  in  column  2. 

TABLE  IV. 

PRESSURES  IN  POUNDS  PER  SQUARE  MILLIMETER  OF  NOZZLE, 
WITH  f-lNOH  TUBING,  ARRANGED  AS  IN  FIG.  16. 


Steam  Pressure, 
Gauge. 

Arrangement 
A. 

Arrangement 
B. 

Arraneement 
C 

Arrangement 

120 
90 
60 
30 

0.4176 
0.3018 
0.1860 
0.0849 

0.3720 
0.2659 
0.1632 
0.0742 

0.3509 
0.2.V79 
0.153? 
0.0733 

0.6805 
0.4'.i94 
0.2998 
0.1414 

Fourth,  the  arrangement  at  D  was  adopted,  with  the  upper  half- 
bend  stationary  and  the  two  lower  ones  mounted  on  the  target,  re- 
sults tabulated  in  column  5. 

In  comparing  results  it  should  be  noted  that,  if  there  were  no 
losses  through  friction,  or  otherwise,  the  pressures  realized  with 
arrangements  A,  B,  and  C,  Fig  16,  should  be  the  same,  since 
the  tubes  all  turn  the  jet  through  180  degrees  at  the  point  of  dis- 
charge; and  the  pressures  should  also  be  double  those  obtained 
with  the  same  nozzle  discharging  against  the  parabolic  target,  as 
given  in  Table  I.,  where  the  jet  is  turned  through  an  angle  of 
only  90  degrees.  With  the  arrangement  at  D,  where  the  steam  is 
twice  turned  through  180  degrees,  and  discharges  twice  at  this 
angle,  the  pressures  should  be  double  those  in  the  other  cases  with 
the  tubes  and  four  times  those  with  the  parabolic  target. 

Leakage  of  Steam. — In  the  foregoing  tests  where  the  tubing 
was  cut  apart  there  was  a  slight  leak  or  overflow  at  the  first  open- 
ing, when  initial  pressure  was  120  pounds,  and  much  more  at  the 
second  opening.  At  90  and  60  pounds  the  leaking  was  much  less, 
and  at  30  pounds  there  was  practically  no  leak  whatever. 

Another  series  of  tests  was  run  with  a  rectangular  tube  of 
slightly  greater  taper,  hoping  to  have  less  leak  at  openings.  Some 


304 


STEAM  TURBINES 


changes  were  also  made  in  the  nozzle.  With  a  continuous  tube  all 
the  steam  entered  from  the  nozzle  nicely;  but  when  the  three  half- 
turns  were  separated  by  ys2-inch  spaces  all  the  steam  would  not 
enter  and  there  «was  some  waste  at  the  mouth.  There  were  also 


TABLE  V. 

PER  CENT  DROP  IN  PRESSURE,  DUE  TO  OPENINGS 
IN  TUBE. 


Steam 
Pressure, 
•Gauge. 

A-  Inch 

Openings. 

Openings. 

|-Inch 
Openings. 

12) 

6.4 

7.5 

13.6 

90 

3.8 

4.8 

10.0 

60 

2.0 

2.0 

9.3 

leaks  at  each  opening,  but  not  as  much  as  in  the  previous  tests.  At 
90  pounds  there  was  very  little  waste  at  the  openings  and  none  at 
the  mouth.  At  GO  pounds  a  slight  leak  from  second  opening  only 
and  at  30  pounds  no  waste  at  all.  The  effect  was  then  tried  of 
increasing  the  width  of  openings  from  %2  to  %  inch  and  the  pres- 
sures measured  in  each  case.  The  readings  were  first  taken  with 
a  continuous  tube  and  then  the  tube  having  openings  of  %^> 
YIQ  and  y%  inch,  the  results  being  as  in  Table  V. 


CHAPTER  XV 
BODIES  ROTATING  AT  HIGH  SPEED. 

Critical  Speed  of  Rotating  Bodies. — In  the  description  of  the 
De  Laval  turbine,  Chapter  III.,  reference  was  made  to  the  so- 
called  critical  speed  of  the  wheel  and  flexible  shaft  when  rotating 
at  high  velocity.  This  phenomenon  may  be  explained  by  the  aid 
of  the  accompanying  diagrams  : 

In  Fig.  1  is  a  disk  W  mounted  on  a  shaft  A  B  turning  in  ball- 
and-socket  bearings,  as  indicated.  One  side  of  this  disk  is  sup- 
posed to  have  a  dense  section  at  PI,  making  it  heavier  than  the 


Fig.    1.     Disk   and    Flexible   Shaft. 

opposite  side.  The  center  of  gravity  of  the  wheel,  therefore,  will 
lie  to  one  side  of  the  shaft  A  B,  say  on  the  axis  C  D.  Now  if  this 
shaft  and  disk  be  rotated,  the  centrifugal  force  generated  by  the 
heavier  side  will  be  greater  than  that  generated  by  the  lighter  side 
diametrically  opposite  to  it,  and  the  shaft  will  deflect  toward  the 
heavy  side,  as  in  Fig.  2,  causing  the  center  of  the  disk  to  describe 
a  small  circle,  indicated  by  the  dotted  line  at  a.  To  locate  the 
point  at  which  a  weight  should  be  added,  or  on  the  other  hand,  at 
which  metal  should  be  drilled  out  in  order  to  bring  the  piece  into 
balance,  a  piece  of  chalk  is  held  so  that  the  high  side  of  the  disk 
will  just  touch  it  as  it  comes  around.  The  weight  necessary  to  bal- 
ance, to  be  told  by  trial,  is  then  added  opposite  to  the  high  side 
where  the  mark  appears ;  or  else,  if  the  balancing  is  to  be  done  by 
drilling,  metal  is  removed  on  the  same  side  with  the  mark.  In  the 
most  accurate  balancing  it  is  advisable  to  use  a  steel  point  held 
rigidly,  but  which  can  be  fed  up  gradually  until  the  point  makes  a 
faint  scratch  on  the  edge  of  the  disk. 


306 


STEAM  TURBINES 


The  foregoing  conditions  hold  until  a  comparatively  high  speed 
is  reached,  depending  upon  the  weight  of  the  disk  and  flexibility 
of  the  shaft.  A  point  will  eventually  be  reached,  however,  at 


Fig.   2.     Rotation  about   Geometrical  Axis. 

several  thousand  revolutions  a  minute,  when  there  will  momen- 
tarily be  excessive  vibration,  and  then  the  parts  will  run  quietly 
again.  The  speed  at  which  this  occurs  is  called  the  critical  speed 
of  the  wheel,  and  the  phenomenon  itself  is  called  the  settling  of 
the  wheel.  The  explanation  is  that  at  this  speed  the  axis  of  rota- 
tion changes  and  the  wheel  and  shaft,  instead  of  rotating  about 
their  geometrical  center,  begin  to  rotate  about  an  axis  through 


— D 


Fig.    3.     Rotation    about   Axis    of    Gravity. 

their  center  of  gravity,  or  about  the  axis  C  D  in  Fig.  1.  This  is 
illustrated  in  Fig.  3,  where  the  wheel  and  shaft  have  taken  a  new 
position  in  which  the  axis  C  D,  if  extended,  would  pass  through 
the  centers  of  the  two  bearings,  while  the  shaft  is  deflected  so  that 
it  traces  a  circle  shown  by  the  dotted  line  b  in  Fig.  3.  It  is  to  be 
noted,  however,  that  this  circle  is  now  on  the  H,  or  heavy,  side  of 
the  disk  instead  of  on  the  other  side  as  before,  so  that  now  if  one 
were  trying  to  locate  the  point  where  weight  should  be  added  in 
order  to  balance  the  disk,  he  would  find  that  the  chalk  mark  came 
on  the  light  side  of  the  disk,  and  that  the  weight  should  be  added 
on  the  same  side. 


BODIES  ROTATING  AT  HIGH  SPEED  307 

Point  Where  the  Settling  Occurs. — Mr.  Konrad  Anderson 
states  that  the  settling  of  a  rotating  body  occurs  when  the  number 
of  revolutions  is  equal  to  the  number  of  vibrations  which  the 
shaft  makes  with  the  wheel  mounted  upon  it.  That  is,  a  shaft  and 
wheel  have  a  certain  time  of  vibration,  just  as  does  a  pendulum  or 
a  spring,  and  when  this  synchronizes  with  the  time  of  rotation  the 
change  is  supposed  to  occur.  In  the  De  Laval  turbine  the  flexible 
shaft  and  wheel  are  in  such  proportions  that  the  settling  takes 
place  very  quickly,  and  the  critical  speed  is  from  %  to  ^  of  the 
normal  number  of  the  revolutions  of  the  wheel. 

Disks  Supported  by  a  Rigid  Shaft. — In  the  illustrations  thus 
far  shown,  the  body  to  be  balanced  is  represented  as  a  disk  sup- 
ported by  a  flexible  shaft.  While  it  is  only  in  special  cases  that 
the  flexible  shaft  would  be  used,  it  serves  to  illustrate  the  principle 
of  balancing  better  than  if  the  shaft  were  rigid.  If  a  disk  were 
mounted  on  a  rigid  shaft  and  the  rotative  method  of  balancing 
were  to  be  applied,  it  would  be  necessary  to  support  the  shaft  in 
bearings  loosely  connected  to  their  pedestals,  which  would  allow 
the  shaft  and  disk  to  vibrate  freely  under  the  action  of  the  forces 
generated. 

The  disks  of  the  smaller  Curtis  turbines  are  balanced  in  this 
manner.  The  shaft  is  placed  in  a  vertical  position  and  suspended 
from  the  end  of  a  cable  which  is  given  a  rotary  motion  of  the 
desired  speed.  The  lower  end  of  the  shaft  is  steadied  by  a  bear- 
ing held  by  springs,  so  that  it  is  free  to  move  in  a  sidewise  direc- 
tion under  the  influence  of  the  rotating  shaft.  Both  ends  of  the 
turbine  shaft  are  thus  flexibly  supported  and  as  the  disk  rotates  the 
high  side  is  marked  with  chalk  or  blue  pencil  held  close  to  the 
rotating  disk.  The  motion  is  then  stopped,  a  small  piece  of  wire 
twisted  into  the  blades  of  the  disk  at  the  proper  point  for  the  pur- 
pose of  bringing  it  into  balance,  and  the  operation  repeated. 
When  perfect  balance  is  obtained,  the  pieces  of  wire  show  where 
metal  must  be  added  to  or  taken  from  the  disk  to  complete  the 
work. 

Static  Balancing  may  be  successfully  used  for  disks,  if  the 
apparatus  is  sufficiently  delicate.  Fig.  4  shows  a  machine  used  by 
the  De  Laval  Steam  Turbine  Company  for  this  purpose.  The 
piece  marked  A  is  a  coupling  flange  mounted  on  a  vertical  arbor 


308 


STEAM  TURBINES 


ready  for  balancing.  The  machine  is  placed  under  a  drill  press 
and  by  its  aid  the  heavy  side  of  the  casting  is  located  and  enough 
metal  drilled  out  to  bring  the  flange  into  balance.  On  top  of  the 
stand  are  knife  edges  which  carry  a  table  C  with  a  movable  cross 
slide  B.  This  cross  slide  is  fitted  with  a  pendulum  in  the  form  of 


Fig.   4.     Apparatus   for   Static   Balancing. 

a  screw  which  runs  down  into  the  base  and  has  a  weight  at  its 
lower  end.  Half  way  up  there  is  a  pointer  and  a  graduated  scale 
for  indicating  the  position  of  the  pendulum.  The  arbor  for  sup- 
porting the  piece  to  be  balanced  is  at  the  top  of  the  slide  B.  By 
adjusting  the  slide  one  way  or  the  other  the  indicator  is  brought 
exactly  at  the  center  of  the  scale  and  then  the  coupling  is  turned 


BODIES  ROTATING  AT  HIGH  SPEED 


309 


around  half  way  by  hand  without  moving  the  slide.  If  the  indi- 
cator does  not  move,  it  shows  that  the  coupling  is  in  balance  at  this 
point;  if  it  does  move,  the  coupling  must  be  out  of  balance  and 
the  necessary  drilling  is  done.  From  eight  to  a  dozen  points 
around  the  circumference  of  the  coupling  are  tested  in  this  man- 
ner. In  order  to  steady  the  table  and  avoid  wear  of  the  knife 
edges  when  drilling,  there  are  pins  D  which  are  raised  against  the 
bottom  of  the  table  by  a  cam  and  thus  take  the  strain. 

Balancing  Cylinders. — In  attempting  to  balance  a  cylinder,  like 
Fig.  5,  a  heavy  portion  might  come  at  one  end,  as  at  H,  and  per- 
haps at  the  other  end,  also,  but  at  a  different  point  of  the  circum- 
ference, as  at  //*,  so  that  a  considerable  twisting  moment  would 


-A — 


Fig.  5. 


be  introduced.  It  is  desirable,  if  possible,  to  divide  the  cylinder 
into  a  number  of  disks,  as  A,  B,  C,  D,  and  balance  each  one 
separately.  But  this  cannot  always  be  done,  as  in  the  case  of  tur- 
bine drums  and  the  rotors  of  electric  generators.  The  only  way 
with  such  parts  is  to  mount  them  in  loose  bearings  supported  by 
springs  and  then  run  them  by  motor  or  other  means  up  to  the 
required  speed. 

Locating  the  Heavy  Side. — It  is  not  always  easy  to  tell  by  this 
means  where  the  heavy  spot  is  located,  since,  as  explained  in  con- 
nection with  the  flexible  shaft  and  critical  speed,  it  may  under  cer- 
tain conditions  be  on  the  same  side  with  the  high  spot  and  under 
other  conditions  on  the  opposite  side;  and  frequently,  if  the 
cylinder  is  approaching  the  critical  point  the  heavy  spot  will  lie 
somewhere  between  these  two  extremes.  In  the  American 
Machinist  for  February  22,  1906,  E.  R.  Douglas  gives  sugges- 
tions for  finding  the  position  of  the  heavy  spot.  After  rotating 
the  cylinder  up  to  speed  and  marking  the  high  spot  on  each  end 


310  STEAM  TURBINES 

with  chalk,  run  it  in  the  opposite  direction  and  make  similar  marks. 
If  they  are  in  a  different  position  from  the  first  ones  the  heavy 
spot  will  lie  half  way  between  the  two  marks,  but  on  which  side 
can  be  told  only  by  trial.  Attach  heavy  balance  weights  at  each 
end  of  the  cylinder  midway  between  the  first  and  second  marks. 
The  weights  should  be  heavy  enough  to  completely  outweigh  the 
heavy  spots.  Now  it  will  be  evident  that  if  the  weights  are  in  a 
position  coinciding  with  the  heavy  spots,  chalk  marks  made  on 
the  circumference  as  the  cylinder  is  rotated  will  agree  with  the 
marks  previously  made  when  the  cylinder  was  rotated  in  that 
direction ;  but  if  the  weights  are  opposite  the  heavy  spots,  and  are 
heavy  enough  to  overpower  the  latter,  then  the  new  chalk  marks 
will  be  opposite  to  the  original  ones,  indicating  that  the  balance 
weights  should  be  attached  in  the  position  that  the  heavy  ones 
now  occupy. 

Stresses  in  Rotating  Bodies. 

The  calculation  of  the  stresses  in  rotating  bodies  is  often  a  com- 
plex problem.  Fortunately  the  excellence  of  material  now  avail- 
able makes  well  constructed  turbine  parts  safe  against  bursting  at 
the  speeds  that  compound  turbines  usually  run.  In  practice  a  factor 
of  safety  is  needed,  so  that  approximate  methods  may  be  used  in 
calculations,  if  the  factor  is  on  the  safe  side.  If  we  are  dealing 
with  a  disk  in  which  there  are  both  tangential  and  radial  stresses 
(Fig.  6),  we  might  neglect  the  considerable  effect  of  the  radial 
tension  and  suppose  the  disk  to  be  made  up  of  a  series  of  concen- 
tric rings,  in  which  the  tangential  or  hoop  tension  only  would  be 
considered.  Such  a  method,  although  only  roughly  approximate, 
would  be  safe,  since  there  would  actually  be  the  additional  radial 
tension  to  help.  This  method  would  be  preferable  to  calculating 
the  average  tensile  stress  across  the  whole  cross  section  of  a  disk, 
as  is  sometimes  done,  since  the  strength  of  a  disk  is  at  its 
weakest  part  and  it  would  give  way  at  the  point  where  tfye  tension 
was  the  greatest.  (See  Weisbach's  Theoretical  Mechanics,  page 
618,  for  method  of  calculating  the  average  stress.) 

Stresses  in  a  Rotating  Ring. — When  a  cylindrical  ring  that  is 
comparatively  thin  radially,  like  the  rim  of  a  flywheel,  is  rotated 


BODIES  ROTATING  AT  HIGH  SPEED 


311 


about  its  axis,  tension  is  set  up  proportional  to  the  weight  of  the 
material  and  the  square  of  the  circumferential  velocity.     This 
tension  is  due  to  centrifugal  force. 
The  formula  for  the  stress  is 


12  wv 


where     /=stress  in  ring,  pounds  per  square  inch. 
w=weight  of  1  cubic  inch  of  material. 
z/=velocity  of  ring  in  feet  per  second. 
£=32.2. 

Example.  —  At  a  speed  of  400  feet  per  second  what  tension 
would  be  set  up  in  a  ring,  one  cubic  inch  of  which  weighs 
0.28  pound? 

12X0.28X160,000 


_ 


32.2 


=16,700  pounds  per  square  inch,  or  about 
the  bursting  strength  of  cast  iron. 


Fig.  6. 


Stresses  in  a  Rotating  Disk.*  —  An  exact  analysis  of  this  prob- 
lem is  given  by  Stodola  in  his  work  on  steam  turbines.  In  the  gen- 
eral case  in  which  the  disk  is  of  varying  width,  the  equations  giv- 
ing the  stresses  are  complicated,  but  for  the  case  of  uniform 

•Prepared  at  the  request  of  the  author  by  Prof.  G.  A.  Goodenough,  of  the  University 
of  Illinois. 


312  STEAM  TURBINES 

thickness  the  equations   reduce  to  comparatively  simple   forms. 
Without  any  attempt  at  derivation  we  give  the  final  results. 

Let    ze;=weight  of  material  per  cubic  inch. 

N=speed,  R.  P.  M. 

rx  Dinner  radius  in'  inches. 

r2=outer  radius  in  inches. 

£=a  constant  (Poisson's  ratio)  usually  taken  as  0.3  for 
iron  or  steel. 

7-  =  radius  of  any  point  at  which  stress  is  desired. 

±>t=  tangential  stress,  Ib.  per  sq.  inch. 

v$r=radial  stress  (See  Fig.) 
Then 

£  =0.00000355  w  N 


For     r=r2,  that  is,  at  the  outer  circumference, 
St=0.00000355  [(3+£)  (%r?+rf)  -  (l  +  3£)r88]. 

while  at  the  inner  circumference  where  r  =  r1} 

St=0.00000355[(3+£)  (r?+%rf)  -  (1+3*)  r^]. 

It  will  be  observed  that  the  thickness  of  the  disk  has  no  in- 
fluence on  the  stress.  The  factor  outside  the  brackets  depends 
only  in  the  material  of  the  disk  and  the  speed:  that  inside  the 
brackets  upon  the  dimensions  of  the  wheel.  Evidently  the  stresses 
increase  as  the  square  of  the  speed. 

Example.  —  Take  a  flat  disk  40  inches  outside  and  4  inches 
inside  diameter,  material  weighing  0.28  pound  per  cubic  inch  and 
running  at  1,000  R.  P.  M.  Taking  £=0.3  the  tensile  stress  at  the 
inner  circumference  is 

.00000355X0.28X(1,000)2  [3.3  (2X202+22)—  1.9X22] 

=2,600  pounds  per  sq.  in.  approx., 
and  at  the  outer  circumference  it  is 

.00000355X0.28X  (1,000)  2  [3.3  (202+2X22)—  1.9X202] 
=580  pounds  per  sq.  in.  nearly. 


BODIES  ROTATING  AT  HIGH  SPEED  313 

Of  course,  the  larger  stress  is  taken  as  determining  the  strength 
of  the  disk.  From  the  form  of  the  general  equation  it  is  seen  that 
the  maximum  tangential  stress  will  always  be  at  the  inner  cir- 
cumference. 

The  circumferential  speed  in  the  example  above  is  about  175 
feet  per  second.  If  the  disk  were  run  at  five  times  this  speed, 
which  is  something  like  that  of  a  De  Laval  turbine  disk  of  this 
diameter,  the  stresses  will  be  twenty-five  times  as  large — 65,000 
pounds  per  square  inch  for  the  inner  circumference  and  14,500 
pounds  per  square  inch  at  the  outer  circumference.  In  the  De 
Laval  turbine,  however,  the  disk  is  not  made  of  uniform  thick- 
ness, and  the  design  is  such  that  the  tangential  stresses  are  the 
same  at  all  radii. 


CHAPTER  XVI 

NOTES  ON  EFFICIENCY  AND  DESIGN. 

Efficiency  of  a  Turbine.  —  On  page  181  is  an  explanation  of  the 
thermal  efficiency  of  turbines,  useful  in  comparing  the  perform- 
ance of  different  turbines.  In  estimating  the  losses  in  a  turbine 
another  method  for  computing  efficiency  is  used,  which  gives 
entirely  different  and  much  higher  results.  It  consists  in  cal- 
culating the  rate  of  steam  consumption  for  an  ideal  turbine,  in 
which  there  are  assumed  to  be  no  losses  of  any  kind,  and  then 
rinding  the  ratio  of  this  to  the  rate  of  steam  consumption  for  an 
actual  turbine. 

In  an  ideal  turbine  steam  would  expand  adiabatically  from  the 
initial  to  the  final  pressure  and  the  energy  can  be  calculated  by 
any  one  of  formulas  (2),  (7)  or  (8)  for  saturated  steam;  or 
(11)  and  (12)  for  superheated  steam,  Chapter  XIII. 

For  example,  in  (8)  the  foot  pounds  of  energy  per  pound  of 
steam  are  represented  by 

'/;«>i  -02)-<72]  (1) 


One  horse-power  is  equivalent  to  33,000  foot  pounds  per  min- 
ute, or  33,000  X  60  foot  pounds  per  hour.  Hence,  the  number  of 
pounds  of  steam  per  horse-power  per  hour  required  by  the  ideal 
turbine,  in  which  steam  expands  adiabatically,  is 


33000X60 
,  or 


2544.98 


Example.  —  Steam  expands  adiabatically  in  a  turbine  from  165 
pounds  absolute  to  1  pound  absolute.  The  rate  of  steam  consump- 
tion is 

2,544.98 

1,193.0  —  562.69(1.5581  —  0.1329)  —  70 


EFFICIENCY  AND  DESIGN  315 

which  reduces  to  the  fraction 
2,544.98 


321.65 


=  7.912  pounds. 


Now  suppose  an  actual  turbine  to  consume  12  pounds  of  steam 
per  brake  horse-power  per  hour  at  normal  load.  Its  efficiency 
would  be  7.912-^12  =  0.659  =  65.9  per  cent. 

Analyzing  Turbine  Losses. — By  the  foregoing  method  the  total 
losses  in  a  turbine  can  be  determined;  but  it  is  a  difficult  matter 
to  subdivide  these,  finding  what  per  cent  is  due  to  one  cause  and 
what  to  another. 

It  has  been  shown  by  H.  F.  Schmidt*  that  a  step  in  this  direc- 
tion can  be  taken  for  a  given  turbine,  by  plotting  the  total  steam 
consumption  curve  for  the  turbine.  It  is  well  known  in  engine 
practice  that  this  curve,  when  plotted  for  a  throttling  engine, 
becomes  a  straight  line,  while  for  an  automatic  engine  it  is 
usually  a  curved  line.  In  the  case  of  the  turbine,  it  is  approxi- 
mately a  straight  line,  without  regard  to  type. 

In  Fig.  1  the  ordinates  of  the  chart  represent  pounds  of  steam 
per  hour  and  the  abscissas  electrical  horse-power.  Points  were 
plotted  for  the  weight  of  steam  used  per  hour  by  a  500  Kw.  tur- 
bine at  different  loads  and  the  line  AD  drawn  through  them, 
which  is  the  total  steam  consumption  line  for  the  turbine.  This 
line  cuts  the  vertical  coordinate  at  a  point  above  the  base  line  and 
intersects  the  latter  to  the  left  of  the  zero  point. 

A  similar  line  CO  was  then  drawn  for  an  ideal  turbine  operating 
without  losses  between  the  same  pressures  as  the  actual  turbine. 
Since  there  are  assumed  to  be  no  losses,  the  weight  of  steam  used 
per  hour  is  directly  proportional  to  the  load  and  the  line  for  the 
ideal  turbine  passes  through  the  zero  point  of  the  chart. 

Now,  it  will  be  evident  that  a  certain  amount  of  power  is  re- 
quired to  drive  the  rotor  of  a  turbine,  and  overcome  the  journal 
friction,  the  windage  of  the  generator  fields,  and  the  friction  of 
the  drum  or  disks  rotating  in  the  atmosphere  of  steam  encased 
in  the  turbine.  The  loss  due  to  these  causes  is  constant,  or  nearly 


hNotes  on  the  Steam  Turbine,  Street  Railway  Journal,  June  25,   1904. 


316 


STEAM  TURBINES 


so,  at  all  loads,  while  another  class  of  losses,  such  as  leakage, 
nozzle  losses,  etc.,  is  directly  proportional  to  the  load.  At  no 
load,  or  where  the  power  delivered  by  the  turbine  becomes  zero, 
the  variable  losses  disappear  while  the  constant  losses  become  the 
load  of  the  turbine.* 

Hence  0£>  on  the  diagram  represents  the  power  absorbed  by  the 
constant  losses  at  zero  load,  and  therefore  at  all  loads.    BO,  drawn 


S;000- 
fjOOO- 
G;000- 
5;000- 
4;000- 
3;000- 
2^000- 


100          0          100        200        300        400        500        600        TOO        800        900       1,000 
Electrical  Horse  Power 

Fig.   1.     Diagram  of  Total   Steam   Consumption    Curves. 


through  0,  parallel  with  AD,  is  the  total  steam,  consumption  line 
for  the  turbine,  with  the  constant  losses  deducted,  and  the  in- 
clination of  CO  to  £0  shows  how  the  variable  losses  increase  with 
the  load. 

The  power  represented  by  the  losses  in  the  turbine  can  be  esti- 
mated from  the  diagram.     Suppose  we  wish  to  find  what  it  is  at 


*As  a  matter  of  fact  the  losses  due  to  leakage,  friction  in  the  nozzles,  etc..  do  not 
entirely  disappear  at  no  load,  when  a  turbine  is  running  light,  but  they  are  very 
small  in  amount  and  in  our  analysis  must  be  classed  with  the  constant  losses. 


EFFICIENCY  AND  DESIGN  317 

• 

640  horse-power.  From  this  point  on  the  base  line,  Fig.  1,  trace 
vertically  to  A,  and  then  horizontally  to  C.  Distance  x  shows  the 
power  that  could  theoretically  be  developed  by  the  steam  if  there 
were  no  losses;  y  the  part  of  this  power  that  goes  into  useful 
work;  z  the  part  that  is  wasted;  w  the  part  absorbed  by  the 
variable  losses ;  and  u  the  part  absorbed  by  the  constant  losses. 

Variable  Losses. — Having  separated  the  constant  from  the 
variable  losses,  the  latter  can  be  divided  into  their  elements  only 
by  a  study  of  the  conditions  and  of  such  tests  upon  nozzles,  vanes, 
etc.,  as  are  available. 

The  variable  losses  are  due  chiefly  to : 

1.  Leakage. 

2.  Radiation. 

3.  Residual  velocity  of  the   steam  leaving  the  last  row   of 
buckets. 

4.  Imperfect  action  of  the  steam  in  the  nozzles  and  blade 
channels,  due  to  friction,  eddying,  etc. 

The  latter  (No.  4)  may  be  placed  under  two  heads,  one  called 
''thermal"  and  the  other  "hydraulic."  The  thermal  losses  are 
caused  by  the  failure  of  the  nozzles  and  blade  channels  to  properly 
expand  the  steam  and  convert  its  heat  energy  into  mechanical 
energy,  while  the  hydraulic  losses  are  caused  by  the  failure  of  the 
moving  vanes  to  convert  all  the  mechanical  energy  of  the  steam 
into  work  at  the  spindle. 

Losses  in  1,250  Kw.  Turbine. — To  illustrate  what  has  preceded, 
take  the  second  test  on  the  1,250  Kw.  Westinghouse-Parsons 
turbine,  page  190.  This  test  was  with  saturated  steam,  with  a 
mean  initial  pressure  of  143  pounds  and  a  condenser  pressure  of 
one  pound  absolute. 

The  theoretical  steam  rate,  calculated  by  formula  (1)  is  7.928 
pounds  per  horse-power  hour. 

The  results  of  the  test  are  given  in  terms  of  electrical  horse- 
power, and  it  will  simplify  the  analysis  to  eliminate  the  generator 
efficiency  from  the  results  and  plot  the  diagram  in  terms  of 
brake  horse-power.  On  page  178  the  efficiencies  of  the  generator 
of  this  turbine  are  given  as  follows:  Full  load,  0.96;  half  load, 
0.93 ;  quarter  load,  0.86,  and  by  using  these  the  brake  horse- 


318 


STEAM  TURBINES 


power  may  be  calculated  from  the  electrical  horse-power.  The 
second  column  below  gives  the  brake  horse-power  for  different 
loads  and  the  first  column,  taken  from  the  results  of  the  test,  the 
total  steam  consumption.  Using  these  values,  the  total  steam  con- 
sumption curve  for  the  turbine  is  then  plotted. 


Full  load, 
Half  load, 
Quarter  load, 


Steam  per  Hour. 

25,639 

19,334 

9,295 


Brake  Horse-Power. 
1,904.5 
1,401.0 
521.7 


In  Fig.  2  AD  is  the  total  steam  consumption  curve,  and  BQ  is 
the  curve  with  constant  losses  deducted. 


500 


500  1,000  1,500  2,000 

Brake  Horse  Power 

Fig.  2.     Diagram  for  1,250  Kw.  Turbine. 


To  locate  CO,  the  curve  for  the  ideal  turbine,  we  have,  steam 
used  per  hour  by  the  ideal  turbine  at  normal  load  =  7.928  X  1,904.5 
(from  table  above)  =  15,095  pounds,  from  which  the  point  cor- 
responding to  normal  load  can  be  located ;  and  this  point,  together 
with  the  zero  point,  will  locate  the  line. 

To  obtain  the  efficiency,  15,095-7-25,639  =  58.9  per  cent. 

The  brake  horse-power  at  normal  load  is  approximately  1,904.5. 
This  is  58.9  per  cent  of  the  total  power  represented  by  the  avail- 


EFFICIENCY  AND  DESIGN  319 

able  energy  in  the  steam.  Hence,  power  that  could  theoretically 
be  produced  =  1,904.5  -f-  0.589  =  3,233.4;  and  3,233.4—1,904.5  = 
1,328.9  =  power  required  to  overcome  losses.  Of  this,  275  horse- 
power (from  the  diagram)  is  required  for  the  constant  losses 
and  1,328.9—275  =  1,053.9  for  the  variable  losses  at  full  load. 

Since  the  efficiency  of  this  turbine  is  58.9  per  cent  the  losses 
amount  to  41.1  per  cent;  and  of  these,  275  horse-power,  or  14.4 
per  cent  (obtained  from  the  chart),  are  required  to  turn  the  rotor, 
leaving  26.7  per  cent  for  distribution  among  the  several  items  that 
make  up  the  variable  losses. 

Unfortunately,  there  are  no  available  data  by  which  the  analy- 
sis can  be  continued  in  the  case  of  the  Parsons  turbine.  We  do 
not  know  the  residual  velocity,  and  it  can  only  be  determined  by 
the  application  of  a  gauge  or  manometer  to  the  last  step  of  the 
turbine,  by  which  the  pressure  can  be  found.  This  velocity  can  be 
assumed  and  the  radiation  assumed,  leaving  the  losses  from  leak- 
age and  imperfect  action  in  the  blade  channels  to  be  obtained  by 
subtraction ;  but  the  results  would  be  but  little  better  than  a  guess 
in  the  absence  of  definite  data. 

Notes  on  Design. 

In  turbines  which  are  divided  into  stages,  it  is  desirable  to  have 
an  equal  amount  of  energy  utilized  in  each  stage,  and  an  impor- 
tant problem  is  to  determine  the  steam  pressures  and  quality  of 
steam  in  the  different  stages  under  these  conditions. 

If  expansion  were  adiabatic  and  the  water  of  condensation  were 
to  settle  out  of  the  steam  in  each  stage,  leaving  the  steam  dry, 
the  number  of  stages  and  the  pressure  at  each  stage  could  be  found 
approximately  by  aid  of  the  charts  at  the  end  of  the  volume. 
Thus,  if  a  turbine  is  to  operate  between  steam  pressures  of  165 
pounds  and  one  pound  absolute  and  the  steam  is  to  have  a  velocity 
of  1,200  feet  per  second  at  each  stage,  we  find  from  diagrams  1  and 
2  that  the  pressure  will  drop  in  the  first  stage  to  117  pounds ;  in 
the  second  to  82  pounds ;  in  the  third  to  57  pounds,  etc.,  approxi- 
mately 13  stages  being  required  in  all. 

This  assumption  in  regard  to  the  dryness  of  the  steam,  however, 
is  probably  not  in  accord  with  the  conditions  and  the  problem  can 


320  STEAM  TURBINES 

be  more  satisfactorily  handled  by  aid  of  the  temperature-entropy 
diagram. 

Temperature-Entropy  Diagram  for  Stage  Turbine. — Assume 
the  initial  and  final  pressures  to  be  165  pounds  and  one  pound, 
respectively,  the  turbine  to  have  three  stages  and  the  steam  to 
expand  adiabatically.  By  formula  (8),  Chapter  XIII.,  the 
kinetic  energy  acquired  per  pound  of  steam  is  found  to  be 
250,251  foot  pounds,  requiring  an  expenditure  of  321.66  heat 
units.  For  the  construction  of  the  temperature-entropy  diagram 
we  have  the  following  data : 

Temperature  saturated  steam  at  165  Ib.  abs.  =  365.88. 

Entropy  of  water  at  165  Ib.  abs.  =  0.523. 

Entropy  of  steam  at  165  Ib.  abs.  =  1.558. 

Temperature  saturated  steam  at  1  Ib.  abs.  =  101.99. 

Entropy  dry  saturated  steam  at  1  Ib.  abs.  =  1.985. 

Points  b  and  c,  Fig.  3,  represent  the  entropy  of  water  and 
steam  respectively  for  the  higher  temperature.  Through  b  draw 
the  straight  line  ba  intersecting  the  32-degree  point  on  the  vertical 
coordinate.  This  is  the  water  line  of  the  diagram  and  closely 
approximates  the  true  water  line,  which  is  a  logarithmic  curve. 
The  entropy  of  point  a  at  a  temperature  101.99  degrees  is  found 
by  measurement  to  be  0.11.  Through  a  draw  the  horizontal  line 
ad,  and  through  c  the  adiabatic  line  cd.  The  line  ce  is  the  sat- 
urated steam  line  and  the  entropy  of  point  <?=  1.985. 

The  area  abed  represents  the  number  of  available  heat  units 
per  pound  of  steam  and  should  equal  321.66,  but  actually  equals 
327.61,  because  the  water  line  was  taken  as  a  straight  instead  of 
a  curved  line. 

We  now  have  the  geometrical  figure  abed  in  the  form  of  a 
trapezoid,  to  be  divided  into  three  equal  parts,  each  part  repre- 
senting the  energy  expended  in  one  stage  of  the  turbine.  This 
is  a  geometrical  problem,  merely,  and  may  be  solved  by  the  fol- 
lowing formula,  taking  the  values  in  terms  of  temperature  and 
entropy  units.* 

Let  A  =  length  of  top  of  trapezoid  in  entropy  units. 
B  •=  length  of  bottom  in  entropy  units. 


'Formula  proposed  by  Ralph   E.   Flanders. 


EFFICIENCY  AND  DESIGN  321 

C=height  in  degrees  F. 

P  =  percentage  of  whole  area  which  is  to  be  included  be- 
tween a  line  horizontal  with  the  base  and  the  base  itself. 

//=height  of  horizontal  line  above  base  in  degrees  F. 
Then: 


C      /  v 

H= (B—^/B-— P(B2— A2)  \ 


(3) 


In  the  example  above  /4  =  1.558— 0.523=1.035;  5  =  1.558  — 
0.11  =  1.448;  C  —  365.88— -101.99  =  263.89;  P  =  %  and  %. 
Substituting  in  (3)  and  solving,  H  =  78.59,  when  P  =  %,  and 
165.49  when  P  =  %. 

Taking  these  values  of  H,  the  horizontal  lines  fg  and  hk  are 
drawn,  dividing  the  diagram  into  three  equal  areas.  The  tem- 
perature represented  by  hk  is  101.99  +  78.59  =  180.58;  and  by  fg 
is  101.99  +  165.49  =  267.48. 

In  Fig.  3  ce  is  the  saturated  steam  line  of  the  diagram  and  the 
amount  of  dry  steam  present  at  the  end  of  adiabatic  expansion  is 

od 

ae 

Diagram  Showing  Re  evaporation. — Under  actual  conditions 
there  would  be  friction  and  eddying  of  the  steam  which  would  re- 
tard the  velocity  of  flow  and  part  of  its  kinetic  energy  would  be 
converted  into  heat,  which  would  reevaporate  some  of  the  moist- 
ure, making  the  steam  drier  than  indicated  by  the  above  ratio. 

Fig.  4  shows  how  this  action  may  be  represented  by  the  heat 
diagram.  Taking  the  same  pressures  as  before,  assume  that 
steam  flows  from  the  higher  to  the  lower  pressure  and  that,  after 
discharging,  its  velocity  is  checked  by  eddying  or  otherwise,  so 
that  20  per  cent  of  its  kinetic  energy  is  converted  into  heat  energy. 

The  available  heat  energy  is  represented  by  area  abed,  con-- 
taining  327.61  heat  units  as  in  Fig.  3. 

We  will  assume  two  different  conditions  of  discharge. 

First,  that  it  is  into  a  space  so  large  or  unconfmed  that  the 
reevaporation  will  have  no  tendency  to  raise  the  steam  pressure. 


322  STEAM  TURBINES 

The  evaporation  of  the  moisture  will  then  take  place  at  constant 
temperature  and  the  change  in  entropy  will  be  along  the  isother- 
mal line  de.  If  this  change  is  dh,  the  area  d'dhti  under  dh  will 
represent  the  work  of  reevaporation,  or  327. 61  -j-5  =  65. 52  heat 
units,  the  reevaporation  being  20  per  cent. 

The  line  dh  is  at  an  absolute  temperature  of  101.99  +  460.7 
=  562.67  degrees  and  dh  will  therefore  have  a  value  of  65.52-f- 
562.69  =  0.116  entropy  unit.  This  added  to  1.558,  the  entropy  of 
point  d,  gives  1.674  as  the  entropy  of  the  steam  at  the  completion 
of  the  reevaporation.  The  dry  steam  present  is  therefore 

ah     1.674 

— = =  0.84,  or  84  per  cent. 

ae     1.985 

In  the  diagram  the  area  afgd  is  equivalent  to  the  area  d'dhh' ', 
which,  deducted  from  abed  leaves  fbcg  as  the  net  heat  energy 
converted  into  useful  work. 

Second:  Another  way  to  regard  the  matter  is  to  assume  the 
steam  to  discharge  from  the  nozzle  into  a  closely  confined  space  so 
that  the  reevaporation  will  raise  the  pressure  and  temperature  of 
the  steam.  If  this  change  is  adiabatic  and  20  per  cent  of  the 
energy  is  expended  in  reevaporation,  it  is  obvious  that  the  tem- 
perature will  rise  until  it  reaches  a  point  g,  where  the  percentage 
of  dry  steam  is  84,  as  before  determined.  To  find  the  quality  of 
the  steam,  first  locate  fg,  cutting  off  20  per  cent  from  the  total 
area  of  the  diagram.  By  formula  (3)  it  is  found  to  be  46.64  de- 
grees above  adf  giving  a  temperature  of  101.99  +  46.64=148.63 
degrees.  Entropy  of  dry,  saturated  steam  at  this  temperature 
(point  I,  Fig.  4)  =  1.841;  entropy  of  steam  discharging  from 
nozzle  remains  constant  at  1.558  ;  dryness  of  steam  =  1.558 -r- 1.841 
=  0.84,  as  found  by  the  other  method. 

A  curved  line  drawn  from  c  to  h  will  give  approximately 
the  quality  of  the  steam  at  any  point. 

Fig.  5  is  a  reproduction  of  Fig.  3,  but  with  the  dotted  lines 
fg,  h'k'  and  ad',  showing  the  net  work  and  the  temperature  of 
each  stage  under  the  assumption  that  expansion  is  adiabatic,  re- 
evaporation  is  20  per  cent  and  that  the  latter  action  is  adiabatic, 
as  above  explained.  Then  : 


EFFICIENCY  AND  DESIGN 


323 


Net  work  in  first  stage  is  fbcg  and  temperature  is  x  instead 
of  x\  net  work  in  second  stage  is  tif'g'k'  and  temperature  is  y' 
instead  of  3; ;  net  work  in  last  stage  is  a'tik'd'  and  final  tempera- 


400 


300 


Sioo 


-4G0.7 


400 


100 


Q 


460.7 


d'liti 


0.4  0.8  1.2  1.6  2.0          0  0.4  0.8  12  1.6 

Entropy  Entropy 


20 


Fig.  3.    Temperature-Entropy  Diagram  for     Fig.  4.    Temperature-Entropy    Diagram    Show- 
Stage  Turbine.  ing  Re-evaporation. 


ture  is  z  instead  of  z,  showing  that  the  heat  given  up  by  friction 
and  eddying  is  carried  along  from  stage  to  stage,  and  discharged 
to  condenser  at  the  end. 

The  temperatures  x ,  y',  and  z  enable  the  quality  of  the  steam 
to  be  found  in  each  stage  by  which  the  area  of  the  passages  may 


324 


STEAM  TURBINES 


be  calculated ;  and  also  the  pressure  differences  existing,  by  which 
the  velocity  of  flow  may  be  calculated. 

Example  in  Design. — Required,  to  proportion  a  500  Kw.  tur- 
bine, multicellular  type,  with  one  wheel  in  each  compartment; 
pressures  165  pounds  and  one  pound  absolute. 

Tests  on  a  500  horse-power  Rateau  turbine  of  this  type  show 
a  brake  efficiency  of  60  per  cent.  Thirteen  per  cent  of  the  power 
at  normal  load  is  required  to  turn  the  rotor,  leaving  27  per  cent 
to  be  distributed  among  the  other  losses. 


400 


200 


IOC 


bl- 


y 


0  0.4  0.8  ].2  1.6  2.0 

Fig.   5.     Diagram    Showing   Re-evaporation. 

The  turbine  will  be  governed  by  throttling  the  steam  and  we 
will  assume  that  the  initial  pressure  is  throttled  to  100  pounds 
absolute  at  normal  load,  leaving  a  small  margin  for  overloads 
without  resort  to  the  secondary  admission  valve. 

There  will  be  10  stages,  this  being  sufficient  to  produce  a  mod- 
erate wheel  velocity  and  to  ensure  that  the  drop  in  pressure  from 
stage  to  stage  will  not  exceed  40  per  cent,  so  that  diverging  nozzles 
will  not  be  necessary. 

The  energy  available  is  found  by  formula  (1)  to  be  289.36  heat 
units,  or  225,122  foot  pounds ;  and  .22,512  per  stage,  giving  a 
velocity  of  discharge  of  1,204  feet  per  second.  Tests  with  con- 
verging nozzles  indicate  that  the  actual  velocity  will  be  within 


EFFICIENCY  AND  DESIGN  325 

2  per  cent  of  the  calculated ;  and  the  actual  velocity  of  discharge 
may  therefore  be  taken  at  1,180  feet  per  second. 

Making  the  wheel  vanes  symmetrical,  we  find,  formula  (10), 
Chapter  XIV.,  wheel  velocity  w  =  628  feet  per  second;  and  by 
formula  (11)  residual  velocity  of  steam  =  430  feet  per  second. 

To  determine  the  area  of  the  guide  passages  or  nozzles  three 
additional  items  must  be  known:  (1)  the  weight  of  steam  re- 
quired; (2)  the  pressure  at  each  stage;  (3)  the  dryness  of  the 
steam  at  each  stage. 

1.  With  a  turbine  efficiency  of  0.60  and  a  generator  efficiency 
of  0.94,  the  combined  efficiency  will  be  0.564.    To  obtain  weight 
of  steam,  500  Kw.  — 670   H.  P.  and  (670  X  500)  -=-0.564  =  653,- 
369   foot  pounds  that  must  be  provided  for  per  second.     The 
available  energy  per  pound  of  steam  was  found  to  be  225,122  foot 
pounds.     Hence,  weight  of  steam  required  =  653, 369  ->  225,122  = 
2.9  pounds  per  second. 

2.  The  pressure  in  each  stage  is  to  be  determined  by  the  aid 
of  formula  (3),  following  the  approximation  of  either  Fig.  3  or 
Fig.  5,  and  carrying  through  the  calculation  for  the  10  stages. 
This  will  give  the  temperatures  for  each  stage,  from  which  the 
pressures  can  be  obtained. 

3.  The  dryness  of  the  steam  at  the  different  stages  is  to  be 
obtained   from   the   diagram   by   estimating   the   evaporation   as 
already  explained ;  either  doing  this  at  the  end  of  expansion  and 
sketching  in  the  temperature-entropy  line  for  the  mixture  of  the 
steam  and  vapor  (ch  in  Fig.  4),  or  by  calculating  the  entropy  at 
different  points  and  plotting  the  curve. 

To  arrive  at  the  probable  reevaporation,  let  us  assume  the  losses 
as  follows : 

Per  cent. 

1.  Constant  (to  turn  rotor) 13 

2.  Radiation  and  leakage 5 

3.  Residual  velocity  (last  stage) 2 

4.  Friction  in  nozzles 2 

5.  Friction  and  eddying  in  channels,  etc 18 

40 


326  STEAM  TURBINES 

Of  these,  a  part  of  (1),  due  to  friction  of  the  rotating  disks  in 
the  steam,  represents  conversion  of  mechanical  work  into  heat. 
If  the  turbine  is  designed  so  that  steam  blows  directly  from  one 
stage  into  the  next,  only  a  part  of  (5)  will  cause  reevaporation; 
but  if  the  steam  is  brought  to  a  standstill  in  each  stage,  the  whole 
18  per  cent  will  act  in  this  way.  It  probably  will  not  be  far  wrong 
to  take  the  reevaporation  at  20  per  cent. 

It  will  be  realized  that  calculations  based  on  normal  load  become 
null  and  void  at  other  loads  and  that  the  turbine  problem  is  one 
of  compromise  between  different  running  conditions.  The  final 
and  most  important  determination  of  proportions  must  be  em- 
pirical, based  on  tests  of  the  completed  machine. 


CHAPTER  XVII 

THE  COMMERCIAL  ASPECT  OF  THE  TURBINE. 

Limitations  of  the  Turbine. — The  field  of  the  turbine  is  limited 
by  its  relatively  high  speed  and  the  facts  that  it  is  a  one-speed 
machine  and  is  non-reversible,  in  the  sense  that  the  reciprocating 
engine  is  reversible,  without  duplication  of  its  main  running  parts. 
The  high  speed  of  the  turbine  precludes  its  use  for  driving  ma- 
chinery by  belt  or  ropes,  unless  some  form  of  balanced  reduction 
gearing  is  employed,  which  no  mechanic  would  advocate,  except 
for  the  smallest  units.  Because  of  these  limitations  the  turbine  is 
restricted  to  driving  direct-connected  apparatus,  such  as  gen- 
erators, centrifugal  pumps,  blowers,  and  the  propellers  of  ships; 
and  in  the  latter  case  special  expedients  must  be  adopted  to  secure 
the  necessary  speed  variation  and  reversal,  as  explained  in 
Chapter  XX.  The  turbine  is  essentially  a  central  station  machine, 
for  the  generation  of  electric  power. 

The  Field  of  the  Turbine. — The  extent  to  which  the  turbine 
has  laid  hold  of  the  central  station  field  may  not  be  generally 
appreciated,  but  is  shown  in  a  striking  manner  by  the  records  of 
sales  of  the  turbine  manufacturers.  Up  to  April,  1906,  as  recorded 
in  the  National  Electric  Light  Association  Report  upon  steam 
turbines  for  that  year,  the  Westinghouse  Machine  Company  had 
either  installed  or  on  order  turbines  for  180  customers;  and  of 
these  120  were  for  electric  light,  traction  or  power  companies. 
Also,  in  a  list  of  turbines  of  500  Kw.  or  over,  installed  by  the 
General  Electric  Company,  or  ordered  of  them,  about  140  out  of 
175  purchasers  were  electric  light,  traction  or  power  companies. 

A  similar  condition  also  exists  in  England  and  on  the  continent. 
In  the  paper  just  mentioned,  is  a  report  by  W.  C.  L.  Eglin,  who 
traveled  abroad  to  investigate  the  turbine  situation.  He  says: 
"There  can  be  no  doubt  as  to  the  status  of  the  turbine  as  a  prime 
mover,  in  the  generation  of  electrical  energy  by  the  use  of  steam, 
for  all  of  the  recently  designed  stations  which  were  visited  were 
equipped  with  turbines  and  some  of  the  older  stations,  built 
originally  for  engine-driven  units,  are  being  increased  in  capacity 
by  installation  of  turbine  units.  The  only  observed  instance  in 


328  STEAM  TURBINES 

which  this  is  not  true  is  that  of  the  Metropolitan  station  in  Paris, 
which  is  being  completed  by  the  installation  of  one  large  engine- 
driven  unit It  is  rather  interesting  to  note  that  in  interviews 

with  leading  engineers  no  question  was  ever  raised  as  to  the 
comparative  merits  of  engines  and  turbines  for  electric  light 
plants." 

While  the  turbine  in  all  sizes  is  very  successful  for  electric 
generating,  the  distinct  field  that  it  has  won  for  itself  is  that  of 
large  units,  in  central  stations.  Large  steam  engines,  with  their 
heavy  frames,  ponderous  moving  parts  and  large  generators,  are 
in  marked  contrast  to  the  small  and  compact  turbine  units  of 
corresponding  power. 

The  Field  of  the  Reciprocating  Engine. — The  power  for  rolling 
mills,  blast  furnaces,  waterworks,  mine  hoisting,  air  and  ammonia 
compressing,  etc.,  will  be  furnished  by  the  piston  engine  for  a 
long  time  to  come.  Rolling  mills  have  been  driven  by  electricity 
to  a  limited  extent,  and  centrifugal  air  compressors,  turbine 
driven,  have  been  successfully  used.  But  in  general  the  turbine 
must  wait  upon  the  development  of  such  apparatus  before  it  can 
enter  the  above  fields.  It  is  also  safe  to  say  that  the  Corliss  or 
similar  type  of  engine  will  continue  to  be  used  for  mill  work, 
where  driving  by  belt  or  ropes  is  in  vogue. 

The  competition  to  be  met  in  electric  generating  will  depend 
upon  the  future  development  of  the  turbine.  The  design  of  Par- 
sons turbines  is  not  well  adapted  to  small-sized  units,  and  turbines 
of  this  type  are  not  now  built  in  this  country  in  powers  of  less  than 
400  Kw.,  or  about  600  horse-power.  As  long  as  these  turbines 
are  not  made  in  the  smaller  sizes,  the  only  competition  with  re- 
ciprocating engines  of  less  than  600  horse-power  will  be  turbines 
of  the  impulse  type,  such  as  the  De  Laval  and  the  Curtis,  and 
there  is  thus  a  comparatively  clear  field  for  the  several  types  of 
engines  made  in  these  sizes.  Their  real  competitor  at  the  present 
time  is  the  gas  engine  rather  than  the  turbine. 

The  possibility  of  engines  of  intermediate  sizes,  say  from  600  to 
1,500  horse-power,  meeting  turbines  on  an  equal  footing  in  com- 
petition depends  mainly  upon  whether  land  values  and  space  avail- 
able are  at  a  premium.  If  such  is  the  case  the  turbine  would 
naturally  be  selected ;  but  if  not,  many  engineers  would  select  in 


COMMERCIAL  ASPECT  OF  THE  TURBINE  329 

preference  compound  engines  of  medium  speed,  which  are 
economical  and  reliable  and  reasonably  compact. 

Turbine  Advantages,  as  usually  claimed,  are  about  as  follows : 
High  economy  under  variable  loads;  small  floor  space;  uniform 
angular  velocity  and  close  speed  regulation ;  freedom  from  vibra- 
tion; inexpensive  foundations;  ease  of  erection  and  quickness  in 
starting;  steam  economy  not  seriously  impaired  by  wear  or  lack 
of  adjustment;  small  cost  for  maintenance  and  attendance;  but 
little  danger  from  water  entrained  in  the  steam ;  adapted  for  high 
superheat ;  water  of  condensation  free  from  oil. 

Reciprocating  Engine  Advantages. — Rather  than  call  attention 
to  special  features,  engine  builders  point  to  the  proven  reliability 
of  the  reciprocating  engine;  to  the  fact  that  it  is  in  no  sense  an 
experimental  or  undeveloped  device ;  that  its  condensing  system 
is  simple,  requiring  only  a  small  quantity  of  cooling  water;  and 
that  high  economy  is  obtained  without  the  use  of  superheated 
steam. 

Several  of  these  claims  for  the  turbine  and  engine  will  bear 
discussion. 

Comparative  Economy. — This  has  been  considered  in  its  dif- 
ferent phases  in  Chapters  IX.  and  X.  A  point  in  this  connection, 
not  as  generally  appreciated  as  it  should  be,  is  that  a  compound 
engine  will  not  show  up  creditably  under  variable  loads  unless 
properly  designed  and  adjusted.  It  is  held  by  some  authorities  that 
under  variable  loads  the  best  results  cannot  be  obtained  with  the 
drop  cut-off  gear  of  a  Corliss  engine.  With  this  gear  the  initial 
pressure  in  the  cylinder  approximates  the  steam  pipe  pressure  at 
all  points  of  cut-off,  and  it  is  held  that  better  economy  is  to  be 
obtained  by  throttling  the  steam  in  the  high-pressure  cylinder  at 
short  cut-offs.  This  is  easily  accomplished  by  the  use  of  shifting 
eccentrics  and  a  shaft  governor. 

In  the  matter  of  adjustment,  tests  on  one  of  the  7,500  horse- 
power engines  of  the  59th  Street  station  of  the  Interborough 
Rapid  Transit  Company,  New  York,  show  what  effect  this  may 
have.  When  running  with  a  load  equally  divided  between  two 
cylinders  the  steam  rate  at  4,000  Kw.  was  17.4  pounds  and  at 
6,000  Kw.  19  pounds.  Afterwards,  by  adjusting  the  low-pressure 
gear  the  receiver  pressure  was  changed,  with  results  at  these  two 


330 


STEAM  TURBINES 


loads  of  17  and  17.5  pounds  per  Kw.  hour.*  It  is  reasonable  to 
suppose  that  a  large  proportion  of  compound  engines,  even  if 
designed  along  the  lines  of  best  economy,  are  not  running  under 
proper  adjustment,  and  this  fact  must  be  considered  in  connec- 
tion with  the  comparative  economy  of  turbines  and  engines;  for 
there  are  no  adjustments  to  be  made  on  the  turbine  that  can 
seriously  affect  its  rate  of  steam  consumption. 


Fig.   1.     Relative   Floor  Space  for  1,000  Kw.  Engines  and  Turbines. 

Electric  Generating. — In  running  direct-connected  alternators 
in  parallel,  the  turbine  has  the  advantage  of  a  uniform  turning 
moment  and  its  high  speed  produces  a  powerful  regulating  force 
without  the  use  of  a  flywheel.  There  is  no  reciprocating  motion 
to  be  converted  into  synchronous  rotary  motion.  The  regulation 
of  turbines  is  so  close  that  it  has  been  found  possible  to  run  rail- 
way, power  and  lighting  circuits  from  the  same  machine.  Where 
a  turbine  is  installed  in  a  plant  with  piston  engines  or  water 
wheels,  it  tends  to  have  a  steadying  influence  on  the  whole  system, 
owing  to  its  inertia  effect. 

The  Use  of  Oil. — No  cylinder  oil  is  required  for  the  turbine, 
so  that  the  exhaust  may  be  condensed  and  used  over  and  over 
again  in  the  boilers,  provided  precautions  are  taken  to  prevent 

*See  diagram  in  paper  by  Henry  G.  Stott  on  Power  Plant  Economics,  Proc.  Am. 
Inst.  E.  E.,  January,  1906. 


COMMERCIAL  ASPECT  OF  THE  TURBINE 


331 


oil  coming  from  the  condenser  auxiliaries  intermingling  with  the 
condensed  steam.  The  bearings  of  the  turbine  are  the  only  parts 
that  are  oiled,  and  as  the  lubricant  is  circulated  through  the  bear- 
ings, then  collected,  cooled  and  used  over  again,  there  is  but  little 
opportunity  for  loss.  Returns  from  a  large  number  of  users  of 
Westinghouse  turbines  show  that  only  about  one  quarter  gallon 
of  oil  is  required  for  the  bearings  per  kilowatt  per  year,  which 
is  at  the  rate  of  about  five  cents  a  day  for  a  400  Kw.  turbine. 


"j*::^ 


'^-^J^^^ '  - 


Fig.  2.     Comparison  of  5,000  Kw.  Units. 

Noise. — There  is  a  great  difference  in  the  running  qualities  of 
different  turbine  generators,  some  of  them  operating  quietly  and 
others  producing  a  roar  that  is  very  objectionable.  This  has  been 
overcome  to  some  extent  by  designing  the  rotating  fields  so  as  to 
have  a  smooth  exterior,  and  more  recently,  in  the  Parsons  type  of 
turbine,  the  noise  has  been  largely  obviated  by  encasing  the  gen- 
erator and  then  circulating  air  through  the  casing  by  means  of  a 
blower.  The  casing  deadens  the  noise  and  the  circulation  of  the 
air  cools  the  motor  and  enables  heavier  overloads  to  be  carried 
for  longer  periods  of  time. 


332 


STEAM  TURBINES 


oooo 

0000 

oooo 

0000 


Fig.  3.     Plan  and  Elevation  of  500  Kw.  Westinghouse  Turbine. 


Scale  of  Feet 


Fig.  4.     Elevation  and  Plan  of  500  Kw.  High-Speed  Engine. 


COMMERCIAL  ASPECT  OF  THE  TURBINE  333 

Relative  Space  Occupied  by  Engines  and  Turbines. 

Comparison  of  500  Kw.  Units. — Fig.  1  is  a  graphic  comparison 

of  the  floor  space  required  for  horizontal  turbines,  and  vertical 

and  horizontal  cross-compound  Corliss  engines,  the  basis  of  com- 


Fig.  5.     Plan  and  Elevation  of  500  Kw.   Corliss  Engine. 

parison  being  a  1,000  Kw.  unit,  including  the  direct-connected 
generator,  the  engine  cylinders  being  28  and  56  by  48  inches,  95 
revolutions.* 


*In  pamphlet  by  Edw.  H.  Sniffin,  issued  by  The  Westinghouse  Machine  Company. 


334 


STEAM  TURBINES 


Fig.  2  is  a  comparison  of  the  famous  Reynolds  vertical  - 
horizontal  engine  of  5,000  Kw.,  and  a  Curtis  turbine  of  the  same 
power. 

The  most  favorable  case  for  the  steam  engine  is  to  be  had  by 
selecting  high-  and  medium-speed  engines  for  comparison  with 


Fig.  6.     Plan  and  Elevation  of  500  Kw.  Vertical  Engine. 


Scale  cf  Feet 

Fig.  7.     Curtis    Turbine,    of 
500  Kw. 


turbines  of  corresponding  power.  In  Figs.  4,  5  and  6  are  three 
different  styles  of  engines  of  500  Kw.  capacity,  two  of  which  are 
horizontal  machines  and  are  to  be  compared  with  the  500  Kw. 
Westinghouse- Parsons  turbine,  Fig.  3,  while  the  third  is  a  vertical 
machine  comparable  with  the  500  Kw.  Curtis  turbine,  Fig.  7.  All 


COMMERCIAL  ASPECT  OF  THE  TURBINE 


335 


of  the  diagrams,  for  both  vertical  and  horizontal  machines  are 
drawn  to  the  same  scale. 

Fig.  4  is  a  plan  and  elevation  of  a  McEwen  compound  high- 
speed engine  designed  to  run  non-condensing  and  drive  a  400  Kw. 
A.  C.  generator;  but  when  running  condensing  it  is  ample  for  a 
~>00  Kw.  generator.  The  plan  view  shows  the  top  of  the  founda- 
tion, which  is  six  inches  larger  all  around  than  the  engine  bed. 

The  vertical  engine,  Fig.  6,  is  a  cross-compound  Shepherd  en- 
gine, medium  speed. 

TABLE  I. 
DIMENSIONS  OF  HORIZONTAL  PARSONS  TYPE  TURBINES  WITH  GENERATORS. 


Kw 

R.  P.  M. 

Length. 

Width. 

Height. 

Weight,  Ib. 

bo  cycles  2 

or  j  phase. 

Not  over 

6600  volts. 

300 

3,600 

23ft.    4  in. 

4ft.    Oin. 

5ft.    lin. 

36,500 

500 

3,600 

24ft.    6  in. 

4ft.    Oin. 

5  ft.    1  in. 

40,000 

750 

1,800 

27  ft.    0  in. 

5  ft.  10  in. 

5  ft.    9  in. 

65,200 

1,000 

1,800 

29ft.    Oin. 

6  ft.  10  in. 

6ft.    6  in. 

81,500 

1,500 

1,800 

31  ft.  10  in. 

6  ft.  10  in. 

6  ft.    6  in. 

103,000 

2,000 

1,200 

34ft.    9  in 

9ft.    2  in. 

8ft.    2  in. 

138,000 

3500 

900 

35  ft.    9  in. 

10ft.    6  in. 

9ft.    4  in. 

237,000 

5,500 

720 

46ft.    Oin. 

lift.    4  in. 

10ft.    6  in. 

417,000 

25  cycles  2 

or  3  phase. 

Not  over 

bboo  volts. 

300 

1,500 

23ft.    6  in. 

6ft.   4  in. 

5ft.    9  in. 

51,000 

500 

1,500 

24ft.    Gin. 

6ft.    4  in. 

5ft.    9  in. 

56.600 

750 

1,500 

28ft.    Sin. 

6ft.    4  in. 

6ft.    Oin. 

75,500 

1,000 

1,500 

29ft.    9  in. 

7ft.    6  in. 

6  ft.    9  in. 

104,200 

1,500 

1,500 

32ft.    6  in. 

7ft.    6  in. 

7  ft.    0  in. 

131,000 

2,000 

1,500 

38ft.    Oin. 

9ft.    Oin. 

8ft.    Oin. 

166,000 

3,500 

750 

42ft.    5  in. 

11  ft.    8  in. 

10ft.    5  in. 

356,000 

5,500 

750 

46  ft.  10  in. 

13ft.    2  in. 

lift.    6  in. 

460,000 

7,500 

750 

50  ft.  10  in. 

13ft.    Sin. 

lift.    6  in. 

511,000 

Tables  of  Dimensions. — Table  I.  gives  the  approximate  floor 
dimensions  and  height  of  horizontal  turbines  of  the  Parsons  type, 
with  their  generators,  as  made  by  the  Allis-Chalmers  Company. 
In  Table  II.  are  the  over-all  floor  dimensions  of  horizontal,  cross- 
compound  Corliss  engines,  as  built  by  the  same  firm.  These  two 
tables  will  enable  the  reader  to  make  an  approximate  estimate  of 
the  saving  in  space  to  be  effected  by  installing  a  turbine  of  the 
horizontal  type.  To  facilitate  comparison  of  the  two  tables, 
column  four  was  prepared,  giving  the  approximate  generator 
capacity  of  the  different  sizes  of  engines.  These  values  were 


336 


STEAM  TURBINES 


worked  out  by  first  estimating  the  indicated  horse-power  by  the 
rule-of-thumb  method  of  squaring  the  diameter  of  the  low- 
pressure  cylinder  and  dividing  by  two;  and  then  estimating  the 


TABLE    II. 
DIMENSIONS  OF  HORIZONTAL  CROSS-COMPOUND  CORLISS  ENGINES. 


Size  of  Cylinder. 

Approx. 
Capacity 
Generator 
in  Kw. 

A 

B 

C 

D 

Diam. 
H.  P. 

Diam. 
L.  K 

Stroke. 

12 

24 

30 

200 

18ft.    3  in. 

15ft.    7  in. 

22ft.    Oin. 

19  ft.    0  in. 

12 

24 

36 

18ft.    9  in. 

16ft.    8  in. 

23ft.    Oin. 

19ft.    6  in. 

14 

28 

30 

300 

18ft.    4  in. 

15ft.    Sin. 

24ft.    6  in. 

19ft.    Oin. 

14 

28 

36 

18  ft.  10  in. 

16ft.    Sin. 

25  ft.    0  in. 

19ft.    6  in. 

16 

32 

36 

400 

19ft.    7  in. 

17ft.    Oin. 

26ft.    6  in. 

20ft.    Oin. 

16 

32 

42 

22ft.    Sin. 

17ft.    6  in. 

27ft.    6  in. 

20ft.    6  in. 

18 

36 

36 

500 

19ft.    9  in. 

17ft.    5  in. 

27ft.'  Oin. 

20ft.    6  in. 

18 

36 

42 

22ft.    6  in. 

16  ft.    0  in. 

28ft.    6  in. 

21  ft.    6  in. 

18 

36 

48 

25ft.    3  in. 

18ft.    7  in. 

30  ft.    0  in. 

22ft.    Sin. 

20 

40 

36 

500 

20ft.    Oin. 

19ft.    6  in. 

26  ft.    6  in. 

22ft.    6  in. 

20 

40 

42 

22ft.    6  in 

19  ft.  10  in. 

29ft.    Oin. 

23ft.    Oin. 

20 

40 

48 

25ft.    Oin. 

20ft.    5  in. 

31  ft.    0  in. 

23ft.    9  in. 

22 

44 

36 

750 

20ft.    9  in. 

19  ft.  10  in. 

27ft.    6  in. 

23ft.    6  in. 

22 

44 

42 

23ft.    3  in. 

21  ft.    2  in. 

30ft.    Oin. 

24ft.    Sin. 

22 

44 

48 

25ft.    Sin. 

21  ft.    4  in. 

32ft.    6  in. 

25ft.    Oin. 

24 

48 

42 

750 

24ft.    6  in. 

22ft.    Sin. 

30ft.    6  in. 

25  ft.    2  in. 

24 

48 

48 

27ft.    Oin. 

23ft.    Oin. 

33ft.    Oin. 

26ft.    Oin. 

26 

52 

42 

1,000 

24ft.    Sin. 

23ft.    Sin. 

30ft.    6  in. 

26  ft.  10  in. 

26 

52 

48 

27ft.    2  in. 

24ft.    4  in. 

33ft.    Oin. 

27ft.    6  in. 

26 

52 

60 

32ft.    2  in. 

24  ft.  10  in. 

38ft.    Oin. 

28ft.    Oin. 

28 

56 

48 

1,000 

27  ft.    6  in. 

25  ft.  10  in. 

34ft.    6  in. 

29ft.    6  in. 

28 

56 

60 

32ft.    6  in. 

26ft.    4  in. 

39ft.    Om. 

30ft.    Oin. 

30 

60 

48 

1,250 

27  fl.    9  in. 

26ft.    4  in. 

35ft.    Oin. 

30ft.    6  in. 

30 

60 

60 

32ft.    9  in. 

27ft.    4  in. 

40ft.    Oin. 

31  ft.    6  in. 

32 

64 

60 

1,500 

33ft.    Oin. 

28  ft.  10  in. 

41  ft.    0  in. 

33ft.    Oin. 

34 

68 

60 

1,750 

33ft.    Oin. 

29  ft.    1  in. 

43ft.    Oin. 

31  ft.  10  in. 

36 

72 

60 

2,003 

33ft.    6m. 

29ft.    6  in. 

44ft.    Oin. 

33ft.    6  in. 

COMMERCIAL  ASPECT  OF  THE  TURBINE 


337 


net  kilowatt  capacity  as  two  thirds  of  the  indicated  horse-power. 
The  above  rule  for  indicated  horse-power  is  on  the  basis  of  600 
feet  piston  speed.  In  the  table  the  same  values  have  in  some  in- 
stances been  given  to  engines  of  different  sizes.  In  such  cases 
the  value  ascribed  to  the  smaller  engine  is  on  the  basis  of  about 
700  feet  piston  speed. 
Space  Necessary  for  Condensing  Apparatus.  —  Comparative 


k~ 

^^^^N 

cc 

K\\ws^^ 

XNDENSER 

:^>^s^^ 

& 

1 

1         ^, 

3 

1 

\\NN\\\\\\\\\\\\\ 

J 

Fig.  8.     Arrangement   of   Condensing   Apparatus   for    Curtis  Turbine. 

figures  of  space  required  for  power  units  are  of  but  little  value, 
unless  the  room  occupied  by  the  condensing  apparatus  is  also 
taken  into  consideration.  In  the  engine  plant,  which  is  usually 
equipped  with  jet  or  barometric  condensers,  the  percentage  of 
room  is  very  small  and  can  be  easily  estimated,  because  of  the 
simplicity  of  the  apparatus. 

The  condensing  apparatus  for  turbine  plants  is  fully  described 
in  Chapter  XIX.  If  a  surface  condenser  is  used,  it  must  be  at 
least  double  the  size  required  for  a  reciprocating  engine  and  a 


338 


STEAM  TURBINES 


corresponding  increase  in  the  capacity  of  the  circulating  pump 
and  piping.  In  order  to  maintain  a  high  vacuum  the  air  pump 
employed  in  engine  practice  is  usually  replaced  by  two  pieces  of 
apparatus :  the  hot-well  pump,  which  removes  the  water  of  con- 
densation, and  the  dry-air  pump,  which  exhausts  the  air  and  vapor 
from  the  condenser.  The  dry-air  pump  is  frequently  made  in  the 
form  of  a  two-stage  air  pump  driven  by  a  steam  cylinder  with 
Corliss  valve  gear,  making  a  large  and  complicated  piece  of 
apparatus. 


Base  plate  of 
turbine  and  condenser 


Dry  air-pump 


Fig.  9.     Condenser    Arrangement    for    Parsons    Turbine. 

It  is  difficult  to  estimate  offhand  the  space  required  for  turbine 
condensing  outfits,  because  considerations  of  convenience  in  pipe 
connections  often  make  it  advisable  to  widely  separate  the  parts  of 
the  equipment.  Several  diagrams  are  shown,  however,  indicating 
the  relative  floor  space  occupied  by  turbines  and  their  condensing 
plants,  where  the  condenser  and  auxiliaries  are  compactly  ar- 
ranged. 

Condensers  for  Curtis  Turbines. — In  Fig.  8  is  a  layout  of  con- 
denser and  auxiliaries  for  a  Curtis  turbine,  sketched  roughly 
from  a  blue  print  furnished  by  the  Alberger  Condenser  Company. 
The  accumulator  and  pump  in  the  upper  right-hand  corner  are 
for  the  step  bearing  of  the  turbine  and  have  nothing  to  do  with 


COMMERCIAL  ASPECT  OF  THE  TURBINE  339 

the  condenser.  Sometimes  the  condenser  is  built  into  the  turbine 
base,  in  which  case  space  would  be  required  only  for  the  circu- 
lating and  the  hot-well  and  vacuum  pumps.  It  is  not  good  policy 
to  crowd  the  condensing  apparatus  too  closely  about  the  turbine, 
because  the  machinery  will  not  be  accessible  for  inspection  and 
repairs,  and  room  must  always  be  provided  for  removing  the 
condenser  tubes. 

Condensers  for  Horizontal  Turbines. — It  is  the  usual  plan  to 
place  the  condenser  and  part  of  the  auxiliaries  in  the  basement; 
and  in  the  case  of  a  horizontal  turbine  advantage  can  be  taken  of 
the  fact  that  a  foundation  is  not  required  under  its  full  length,  by 


Fig.   10.     Connection    to     Condenser. 


utilizing  part  of  the  space  under  the  turbine  for  the  condenser. 
The  dry-air  pump  is  so  complicated  a  piece  of  apparatus  that  it  is 
better  to  place  this  on  the  turbine  room  floor,  instead  of  in  the 
basement,  to  ensure  its  having  the  same  care  and  attention,  and 
being  kept  in  as  good  condition  as  the  turbine  itself. 

In  Fig.  7,  Chapter  IXX.,  is  one  of  the  most  compact  arrange- 
ments of  condenser  and  pumps  that  can  be  devised.  The  space 
occupied,  including  room  for  removal  of  condenser  tubes,  is  less 
than  twice  the  area  of  the  baseplate  of  the  turbine  and  generator. 

A  plan  that  meets  with  much  favor  is  to  place  the  condenser 
crosswise  of  and  directly  underneath  the  turbine  as  in  Fig.  9, 
herewith.  The  concrete  piers  supporting  the  turbine  are  located 


340 


STEAM  TURBINES 


at  A,  B  and  C,  as  indicated  by  the  dotted  sections.  An  arch 
would  be  sprung  between  B  and  C  and  I-beams  running  from 
A  to  B  would  be  used  to  stiffen  the  baseplate  at  this  point.  If 
the  condenser  is  of  the  downward-flow  type,  receiving  steam  at 
the  top,  it  would  be  placed  directly  under  the  exhaust  nozzle  of 
the  turbine,  as  indicated,  and  the  various  pumps  could  be  grouped 
as  shown. 

If  a  counter-current  condenser  is  used,  receiving  steam  at  the 
bottom,  it  would  be  necessary  to  locate  it  to  the  left,  next  to  pier 
A,  Fig.  9,  using  an  elbow  in  the  exhaust  pipe.  The  location  and 


Fig.  11.     Two  Turbines   served   by   one    Condenser. 

arrangement  are  indicated  in  Fig.  10.  A  similar  arrangement  of 
piping  would  be  adopted  if  it  were  desired  to  set  the  condenser 
parallel  with  the  turbine  and  alongside  of  the  turbine  foundation, 
as  is  often  done.  Fig.  8,  Chapter  XIX.,  shows  a  turbine  with 
condenser  located  underneath. 

Frequently  space  is  economized,  as  well  as  first  cost  of  ap- 
paratus, by  having  one  condenser  serve  two  turbines,  placing  the 
condenser  between  them,  as  in  Fig.  11.  This  makes  one  of  the 
most  compact  arrangements  and  is  satisfactory  for  small  and 
medium-sized  units. 


COMMERCIAL  ASPECT  OF  THE  TURBINE 


341 


From  the  foregoing  it  appears  feasible  to  arrange  the  con- 
densing apparatus  to  come  into  an  area  equal  to  two  to  three 
times  the  space  occupied  by  the  turbine  and  generator,  in  the 
case  of  horizontal  turbines ;  and  an  area  of  four  to  five  times  the 
space  required  for  the  foundation  of  vertical  turbines,  with  due 
regard  for  the  removal  of  condenser  tubes. 

Enlargement  of  Plant. — The  possibilities  of  the  turbine  as  a 
means  for  the  enlargement  of  an  existing  plant,  is  well  illustrated 
by  the  turbine  installation  of  the  B.  F.  Goodrich  Company,  Akron, 
O.  The  plan  of  the  engine  room  is  as  in  the  accompanying  illus- 


0       4       8      12     16 


200  K-W.  Rotary  Converters 

Fig.  12.     Engine   Room   with   Turbine   Additions. 


tration.*  There  were  originally  the  cross-compound  engines  and 
generators  shown,  and  it  was  found  impossible  by  any  arrange- 
ment of  the  machinery  to  increase  the  capacity  of  the  plant  by 
any  more  than  500  Kw.,  without  extending  the  building,  if  the 
same  type  of  units  was  adhered  to.  Instead,  it  was  decided  to 
install  two  turbines,  as  indicated  in  the  plan,  which  doubled  the 
capacity  of  the  plant  without  disturbing  the  old  arrangement.  It 
will  be  evident  that  another  increase  in  the  capacity  may  be  made 
by -replacing  the  150  Kw.  engine  unit  by  two  400  Kw.  turbo 


*  Originally  appeared  in  Engineering  News. 


342  STEAM  TURBINES 

units.  The  power  house  was  originally  laid  out  for  1,650  kilo- 
watts, but  without  enlarging  the  building  or  replacing  the  large 
engines  it  could  be  made  to  accommodate  2,950  kilowatts,  an  in- 
crease of  80  per  cent.  The  present  engine  plant  of  1,150  Kw. 
occupies  2,630  square  feet,  or  44  per  cent  of  the  total.  The  tur- 
bine plant  of  1,150  Kw.  occupies  980  square  feet  or  16  per  cent  of 
the  total.  The  balance  of  the  space  is  occupied  by  other  apparatus. 

Comparative  Cost  of  Turbine  Outfits  and  Their  Maintenance. 

The  cost  of  complete  turbine  and  engine  outfits  is  practically 
the  same  (exclusive  of  land  and  buildings),  and  such  differences 
as  exist  will  be  found  to  be  no  greater  than  often  met  with  in 
the  cost  of  different  engine  equipments  of  the  same  power.  The 
selling  price  of  turbine  outfits  is  governed  by  the  price  of  engine 
outfits  rather  than  by  the  cost  to  manufacture. 

It  will  be  of  interest  to  compare  cost  figures  for  the  apparatus 
of  two  plants  of  equal  size,  one  engine-driven  and  one  turbine- 
driven.  Such  an  itemized  statement  will  show  the  distribution  of 
expense,  illustrating  how  certain  factors  entering  into  the  engine 
costs,  such  as  the  foundations  and  generator,  are  offset  by  other 
items,  such  as  the  turbine  condensing  system.  The  data  will 
also  assist  the  reader  in  making  his  own  preliminary  estimates. 

Example  for  Comparison. — Assume  the  case  of  a  750  Kw. 
turbine  and  generator  operating  with  150  pounds  steam  pressure 
and  100  degrees  superheat.  The  generator  to  be  60-cycle,  3-phase, 
2,300  volts. 

As  the  equivalent  of  this  a  firm  of  engine  builders  have  pro- 
posed to  supply  a  cross-compound  24  and  50  by  42  engine,  100 
revolutions,  operating  with  150  pounds  pressure,  saturated  steam. 
The  engine  is  rated  at  1,200  horse-power,  or  about  ll/2  times  the 
Kw.  capacity  of  the  turbine.  This  is  ample  to  allow  for  the 
losses  in  the  engine  and  generator  when  comparing  the  indicated 
horse-power  of  an  engine  with  the  net  output  in  kilowatts  of  a 
turbo-generator.  The  proportions  of  the  engine  are  generous, 
giving  ample  power  for  overloads.  The  generator  is  to  be  60- 
cycle,  72-pole,  3-phase,  2,300  volts. 


COMMERCIAL  ASPECT  OF  THE  TURBINE  343 

Estimate  for  Turbine  Outfit: 

Turbine  and  Generator. — 750  Kw.,  150  pounds  steam  pressure, 
100  degrees  superheat;  generator,  60-cycle,  3-phase,  2,300  volts. 
Price,  $22,500. 

Exciter. — Engine-driven,  $1,500. 

Surface  Condenser. — Four  square  feet  cooling  surface  allowed 
per  kilowatt,  or  3,000  square  feet;  60  to  70  pounds  cooling  water 
per  pound  steam.  Condenser  equipped  with  12-inch  centrifugal 
circulating  pump,  driven  by  an  8  by  8  engine,  and  an  air  pump, 
steam  driven.  Price,  $5,200. 

Barometric  Condenser. — If  a  condenser  of  this  type  is  preferred 
it  should  be  capable  of  handling  60  times  the  volume  of  con- 
densed steam.  Price  $i,000. 

Erecting  Condenser. — Two  hundred  dollars. 

Foundation. — Of  concrete,  12.5  feet  deep.  At  $6  per  cubic  yard 
(a  mean  value),  the  cost  would  be  $250. 

Superheater. — Estimate  on  the  basis  of  $2  per  boiler  horse- 
power for  50  degrees  superheat  and  $2.60  per  boiler  horse-power 
for  100  degrees  superheat.  One  boiler  horse-power  is  the 
capacity  to  evaporate  34.5  pounds  of  water  from  and  at  212  de- 
grees F.  The  boilers  necessary  to  generate  steam  for  high-grade 
compound  engines  or  turbines  may  therefore  safely  have  less  than 
one-half  the  rated  power  of  engine  or  turbine.  In  the  present 
case  500  horse-power  boiler  capacity  will  meet  the  requirements, 
and  the  superheater  for  100  degrees  superheat  would  cost  $1,300. 

Estimate  for  Engine  Outfit: 

Engine. — Cross-compound,  24  and  50  by  42;  100  revolutions; 
150  pounds  saturated  steam.  Price  $14,600,  or  $20  per  Kw. 

Generator. — Alternating  current ;  60-cycle ;  72-pole ;  3-phase  ; 
2,300  volts.  Price,  $8,950. 

Exciter. — Engine-driven.     Price,  $1,500. 

Surface  Condenser. — Allow  10  pounds  steam  condensed  per 
hour  per  square  foot  cooling  surface,  and  35  pounds  cooling 
water  per  pound  steam.  Direct-acting  pump  with  circulating  and 
air  cylinders.  Price,  $3,000. 

Barometric  Condenser. — If  preferred,  allow  40  to  50  pounds 
water  per  pound  steam.  Centrifugal  pump  steam  driven.  Price, 


344  STEAM  TURBINES 

$2,500.    An  ordinary  jet  condenser  could  be  installed  for  $2,000. 

Foundation. — Concrete,  12.5  feet  deep;  276  cubic  yards,  $6  per 
yard.  Price,  $1,656. 

Erecting  and  Freight  on  outfit  complete,  $2,500. 

Apparatus  in  Common. — The  feed  pumps,  switchboard,  stack 
and  boilers,  superheater  excepted,  would  be  the  same  for  either 
type  of  plant.  Water  tube  boilers,  delivered  and  erected,  cost 
$14.50  per  boiler  horse-power.  The  piping,  exclusive  of  exhaust, 
would  cost  practically  the  same  in  either  case,  and  may  be  esti- 
mated at  $8  per  boiler  horse-power.  The  expense  of  exhaust 
piping  will  depend  upon  the  location  of  condenser,  which,  in  the 
case  of  the  turbine  is  often  connected  directly  to  the  turbine. 
Smaller  water  piping  is  required  for  the  condenser  with  an 
engine  than  with  a  turbine. 

Summary. — Assuming  the  surface  condenser  for  each  type  of 
apparatus  and  tabulating  the  items  that  differ  in  the  two  types, 
we  have: 

Turbine.  Engine. 

Turbine    and   Generator,  $22,500  Engine,  $14,600 

Surface  Condenser,  5,200  Generator,  8,950 

Erecting  Condenser,  200  Surface   Condenser,  3,000 

Foundations,  250  Foundation,  1,656 

Superheater,  1,300  Erecting,  2,500 

$29,450  $30,706 

Ordinarily  the  surface  type  of  condenser  would  not  be  installed 
with  the  engine,  thus  reducing  the  cost  somewhat. 

General  Figures. — Taking  the  above  figures,  with  the  omission 
of  the  superheater  cost,  we  have,  for  a  750  Kw.  turbo-generator, 
surface  condenser,  foundation  and  installation,  $28,150,  or  $37  per 
Kw.  The  engine  figures  are  $40  per  Kw.  To  compare  with  this, 
the  actual  cost  of  two  larger  units  will  be  given.  One  was  a 
1,500  Kw.  turbo-generator,  which,  with  surface  condenser,, 
foundation  and  installation,  cost  $30.20  per  Kw.  The  other  was 
an  1,100  Kw.  cross-compound  engine,  jet  condenser,  foundation, 
and  installation,  costing  $32.40  per  Kw. 

The  cost  of  turbo-generators  per  Kw.  (60  cycles)  ij  approxi- 
mately as  follows:  3,000  Kw.,  $20;  1,500  Kw.,  $24;  750  Kw., 
$30  ;  500  Kw.,  $32. 


COMMERCIAL  ASPECT  OF  THE  TURBINE  345 

Surface  condensers  for  high  vacuum  cost,  with  necessary  aux- 
iliaries, from  $7  to  $10  per  Kw. ;  and  barometric  jet  condensers 
from  $5  to  $6  per  Kw.  Two  additional  examples  of  condenser 
costs  will  be  cited. 

A  1,550  square  foot  surface  condenser  for  a  750  Kw.  engine; 
26-inch  vacuum  (2  square  feet  per  Kw.)  ;  direct-acting  pump 
underneath  condenser  and  centrifugal  circulating  pump  and  en- 
gine at  end  of  condenser.  Price,  f.  o.  b.  factory,  $2,650,  or,  say 
$3.50  per  Kw. 

A  2,000  square  foot  surface  condenser  for  a  500  Kw.  turbine; 
28-inch  vacuum  (4  square  feet  per  Kw.)  ;  Edwards  air  pump; 
centrifugal  circulating  pump  and  engine.  Price,  f.  o.  b.  factory, 
$3,852,  or,  say  $7.50  per  Kw. 

Cost  of  Land  and  Buildings. — No  estimate  of  turbine  costs  is 
of  any  value  whatever  without  taking  into  consideration  the  cost 
of  land  and  buildings.  In  comparing  turbine  and  engine  costs 
the  important  item  to  be  considered  is  the  investment  for  real 
estate  in  the  two  cases.  Outside  of  the  operating  room  the  space 
required  for  the  plant  will  not  be  materially  different,  whichever 
type  of  apparatus  is  installed.  The  turbine  room  of  the  turbine 
plant,  however,  need  be  only  about  one  half  the  size  of  the  same 
room  in  an  engine  plant,  and  knowing  the  cost  of  building  and 
the  value  of  land  per  square  foot  in  any  community,  the  saving 
can  be  very  quickly  arrived  at  in  the  rough.  Thus,  take  the  case 
of  a  plant  with  four  750  Kw.  units.  By  the  aid  of  Tables  I.  and  II. 
and  a  little  figuring  with  pencil  and  paper,  it  will  be  found  that 
an  engine  room  of  7,600  square  feet  and  a  turbine  room  of  half 
this,  or  3,600  square  feet  will  be  ample  for  all  the  power-gen- 
erating machinery.  At  $5  per  square  foot  the  land  saving  would 
be  $18,000  and  the  saving  in  the  building  might  easily  be  as 
great,  depending  entirely  upon  the  style  adopted. 

Cost  of  Maintenance  and  Operation. — In  a  paper  before  the 
American  Institute  of  Electrical  Engineers,  January,  1896,  Henry 
G.  Stott,  superintendent  of  motive  power,  Interborough  Rapid 
Transit  Company,  New  York,  gives  a  careful  analysis  of  power- 
plant  economics.  He  considers  different  types  of  prime  movers, 
including  gas  engines  and  combinations  of  gas  engines  and  tur 
bines  and  of  reciprocating  engines  and  turbines. 


346 


STEAM  TURBINES 


In  the  paper  is  a  tabulation  of  the  relative  values  of  the 
various  items  necessary  in  the  maintenance  and  operation  of 
power  plants,  and  two  columns  of  the  table  are  reproduced  here- 
with. The  first  column  covers  a  plant  with  compound  condensing 
reciprocating  engines  without  superheat  and  is  derived  from  a 
year's  record  of  actual  costs  in  a  plant  with  7,500  horse-power 
Allis  vertical-horizontal  engines.  The  turbine  compared  with  this 
is  a  5,000  Kw.  unit,  which  is  believed  by  Mr.  Stott  to  have  the 

TABLE  III. 

DISTRIBUTION  OF  MAINTENANCE  AND  OPERATION. 
CHARGES  PER  Kw.  HOUR. 


Maintenance. 

Recipro- 
cating 
Engines. 

2.57 
4.61 
0.58 
1.12 

2.26 
1.06 
0.74 
7.15 
0.17 
61.30 
7.14 
6.71 
1.77 
0.30 
2.52 

Steam 
Turbines 

0.51 
4  30 
0.54 
1.12 

2.11 
0.94 
0.74 
6.68 
0.17 
57.30 
0.71 
1.35 
0.35 
0.30 
2  52 

Boiler  room  

Coal  and  ash  handling  apparatus                   

Electrical  apparafus 

Operation. 
Coal  and  ash  handling  labor  

Removal  of  ashes  ..          

Dock  rental 

Boiler  room  labor  

Boiler  room  oil,  waste,  etc.                

Coal  •  

Water  

Engine  room  mechanical  labor 

Waste,  etc  

Electrical  labor         .  . 

Relative  cost  of  maintenance  and  operation.  . 
Relative  investment  in  per  cent  

100.00 
100.00 

79.64 

82.52 

best  record  for  economy  up  to  date.  It  has  a  flatter  steam  rate 
curve  than  the  engine,  shows  practically  as  good  economy  at 
normal  load  with  saturated  steam  and  a  thermal  economy  6.6  per 
cent  better  with  superheated  steam.  The  various  turbine  items 
are  derived  from  actual  costs. 


Turbine  Troubles. 

In  beginning  the  construction  of  steam  turbines,  it  was  inevita- 
ble there  should  be  difficulties  which  could  not  be  foreseen  and 
which  could  be  overcome  only  by  observing  the  machines  in 


COMMERCIAL  ASPECT  OF  THE  TURBINE  347 

operation  after  they  were  erected  in  the  different  power  plants. 
With  the  turbine  has  come  a  whole  series  of  new  engineering 
problems,  apart  from  the  turbine  itself.  The  greatest  of  these 
has  been  the  electric  generator,  which  must  not  only  be  right,  elec- 
trically, but  must  be  free  from  deformation  and  the  consequent 
destruction  of  balance  at  high  speeds.  This  problem  has  still  to 
be  solved  by  many  of  the  electric  companies  not  directly  interested 
in  turbine  construction.  In  1006  the  author  visited  an  engine 
works  where  an  experimental  turbine  had  been  built  and  was 
standing  idle  on  the  erecting  floor,  because  it  had  not  been  possi- 
ble to  secure  a  satisfactory  generator  for  it. 

Other  problems  have  come  from  the  ability  of  the  turbine  to 
utilize  high  vacuums,  which  has  required  a  vast  improvement  in 
condensing  machinery;  and  from  the  employment  of  highly 
superheated  steam,  which  has  made  provision  necessary  in  the 
turbine  and  connections  to  equalize  the  expansion  due  to  high 
temperatures. 

The  best  testimony  upon  the  success  with  which  these  problems 
have  been  worked  out  is  to  be  had  from  turbine  users  themselves. 
Such  testimony  has  been  published  in  the  1905  report  of  the 
standing  committee  of  the  National  Electric  Light  Association, 
for  the  investigation  of  steam  turbines,  and  is  given  in  condensed 
form  herewith : 

Report  of  Committee  of  the  N.  E.  L.  Association. — In  preparing 
the  1905  report,  95  companies  were  written  to  for  information. 
These  companies  were  operating  turbines  aggregating  over  100,- 
000  kilowatts  and  special  efforts  were  made  to  obtain  a  com- 
plete statement  of  troubles  experienced  with  the  machines.  Re- 
plies were  received  from  59  companies,  and  it  is  interesting  to 
note  that  all  seemed  well  satisfied  with  their  turbines  and  very 
few  admitted  having  any  trouble  with  the  turbines  themselves, 
independent  of  the  condensers  and  other  auxiliary  apparatus. 

Of  the  replies  from  users  of  De  Laval  turbines,  one  reported 
some  trouble  due  to  the  design  of  the  brush  holders  of  the 
dynamo,  which  was  easily  remedied,  while  another  had  trouble 
with  the  oiling.  This  was  due  to  the  filling  up  of  the  spiral  oil- 


348  STEAM  TURBINES 

ways,  which  stopped  the  flow  of  oil,  and  was  overcome  by  the 
use  of  special  oil,  carefully  filtered. 

One  company,  operating  Curtis  turbines,  reported  a  shut- 
down caused  by  a  worn  bearing  on  the  tachometer  connected  with 
the  latch  of  the  emergency  stop  valve;  also  slight  trouble  from 
loose  laminations  in  the  armature  of  the  generator.  On  another 
Curtis  turbine  the  needle  valves  and  the  main  nozzle  valves  were 
warped  by  the  high  degree  of  superheat  employed.  On  still 
another  turbine  of  this  type  there  was  air  leakage  in  the  turbine 
and  also  trouble  from  water  mixing  with  the  oil  lubricating  the 
step  bearing. 

Several  companies  having  Parsons  turbines  reported  difficulties. 
One  experienced  trouble  from  the  oil  solidifying  into  a  jelly  in 
some  of  the  bearings.  Another  reported  the  repeated  cutting 
out  of  the  throttle  valve  seat,  which  was  attributed  to  the  peculiar 
water  used  in  the  locality;  also  excessive  vibration  caused  by  the 
expansion  of  the  exhaust  piping,  throwing  the  turbine  out  of  line. 
This  latter  was  corrected  by  installing  an  expansion  joint  in  the 
exhaust  pipe.  The  field  coils  of  the  generator  of  two  Parsons 
turbines  burned  out  at  less  than  normal  load,  indicating  some 
error  in  design.  There  was  also  a  breakage  of  one  or  more  of  the 
brass  sleeves  of  the  main  bearings  of  these  turbines  and  special 
attention  was  required  to  keep  the  lubricating  system  in  good 
order. 

Three  companies  reported  breakages  of  blades  in  Parsons  tur- 
bines. In  one  case  where  superheated  steam  was  used,  sufficient 
time  was  not  allowed  in  starting  to  warm  up  the  machine  and 
maintain  the  proper  clearance  between  the  blade  tips  and  the 
casing.  In  another  case,  where  there  was  said  to  have  been  no 
rubbing  of -the  blades  against  the  casing,  many  of  the  blades  of 
the  rotor  were  broken  while  running.  Whatever  the  cause  of 
this  may  have  been,  such  accidents  are  now  guarded  against  by 
the  use  of  steel  lacing  to  stiffen  the  blades.  The  third  company 
had  a  few  blades  broken  by  some  foreign  substance  carried  into 
the  turbine  through  the  steam  pipe.  This  did  not  affect  the 
operation  of  the  turbine,  which  continued  to  run  until  it  was 
convenient  to  repair  it. 

It  will  be  seen  from  the  above  that  the  accidents  reported  are 


COMMERCIAL  ASPECT  OF  THE  TURBINE  349 

all  of  a  minor  character,  with  the  exception  of  the  blade  failures. 
Even  when  making  allowance  for  the  reticence  of  the  firms  in- 
terrogated, the  difficulties  experienced  must  be  admitted  to  be  very 
few  and  comparatively  insignificant. 

Danger  from  Water. — It  is  claimed  for  the  turbine  that  it  is 
not  injured  by  water  coming  over  from  the  boiler  in  case  of  exces- 
sive priming.  Instances  are  on  record  where  a  slug  of  water  has 
suddenly  entered  a  turbine,  bringing  the  rotating  member  almost 
to  a  standstill,  without  injury  to  the  machine.  Destruction  of  the 
blading  has  also  occurred  from  this  cause  in  some  cases,  but  this 
is  practically  impossible  except  in  machines  like  the  Rateau  or 
certain  types  of  Parsons  turbines  in  which  the  outer  ends  of  the 
blades  are  unsupported ;  and  even  in  these  machines  the  high- 
pressure  blades  are  so  short  that  damage  from  water  mingled  with 
the  incoming  steam  seldom  results.  Breakage  would  be  more 
likely  to  occur  if  water  should  set  back  from  the  condenser 
through  some  difficulty  with  the  pumps.  At  the  low-pressure 
end  of  the  turbine  the  blades  are  long  and  slender,  and  running  as 
they  do  at  very  high  speed,  sudden  contact  with  the  water  might 
strip  off  the  last  row.  Further  damage  seems  to  be  prevented, 
however,  by  the  next  row  of  fixed  guide  vanes,  which  divide  the 
water  into  small  streams  and  thus  protect  the  other  rotating  mem- 
bers. Compared  with  the  breakages  that  so  often  occur  from 
water  in  a  steam  engine  cylinder,  turbine  troubles  from  water 
seem  insignificant. 

Distortion  of  Casing. — In  Parsons  turbines  there  has  been 
trouble  from  the  distortion  of  the  casing,  resulting  in  the  moving 
and  stationary  parts  coming  in  contact,  tearing  out  some  of  the 
blading.  Trouble  has  been  experienced  from  the  casing  arching 
upward  under  the  effect  of  superheated  steam,  on  account  of 
the  top  of  the  cylinder  expanding  more  than  the  bottom,  this 
being  due  to  the  fact  that  the  shell  was  not  made  symmetrical, 
sometimes  having  ribs  or  heavier  parts  at  the  bottom  than  at  the 
top.  This  has  been  remedied  in  later  machines  by  more  care  in 
the  design. 

Another  cause  for  distortion  has  been  the  "pull"  of  the  con- 
denser at  the  low-pressure  end,  due  to  the  vacuum.  Inasmuch  as 
a  high  vacuum  is  carried,  at  which  pressure  steam  has  a  high 


350  STEAM  TURBINES 

specific  volume,  the  opening  to  the  condenser  is  necessarily  of 
unusually  large  dimensions  and  atmospheric  pressure  distributed 
over  this  opening  produces  a  heavy  stress.  Under  the  most  ap- 
proved form  of  construction  the  exhaust  nozzle  leading  from  the 
turbine  passes  down  through  the  pedestal  at  the  low-pressure  end 
of  the  turbine,  thus  placing  this  stress  directly  upon  the  founda- 
tion in  so  far  as  possible. 

A  corrugated  copper  expansion  joint  is  also  placed  in  the  ex- 
haust piping  just  below  the  turbine  outlet,  to  compensate  for 
unequal  expansion  and  for  any  change  in  the  relative  posi- 
tions of  condenser  and  turbine,  due  to  settling  of  foundations. 

Another  method  that  has  been  tried  consists  in  bolting  the  con- 
denser flange  rigidly  to  the  casing,  with  the  condenser  under  the 
casing.  The  base  of  the  condenser  rests  on  a  flexible  foundation 
of  springs,  sufficient  to  relieve  the  turbine  of  the  weight  of  the 
condenser,  but  allowing  it  to  go  and  come  with  the  turbine.  The 
condenser  is  thus  to  all  intents  and  purposes  a  part  of  the  tur- 
bine, supported  by  the  turbine  foundation,  and  it  will  be  evident 
that  the  "pull"  of  the  vacuum  will  have  no  more  tendency  to  dis- 
tort the  casing  than  will  the  pressure  at  an}'  other  part  of  the 
casing. 

Stripping  the  Blades. — The  most  serious  accident  that  can 
befall  a  turbine  is  that  mentioned  under  the  last  heading,  of  the 
rotating  and  stationary  parts  coming  together  and  the  stripping 
of  the  blades.  This  trouble  has  been  experienced  to  a  greater  or 
less  extent  in  turbines  in  which  the  blades  have  no  protection  or 
support  at  their  outer  ends.  In  the  early  days  of  the  Parsons  tur- 
bine there  was  an  endless  amount  of  trouble  from  blade  failures. 
The  blades  broke,  not  only  when  the  rotating  and  stationary  parts 
came  in  contact,  but  from  no  visible  cause,  one  theory  being  that 
rapid  vibration  of  the  rotating  member  produced  repeated  stresses 
in  the  blades,  leading  to  their  rupture. 

Lately  a  great  deal  of  attention  has  been  given  to  the  blade 
question.  Where  the  material  is  an  alloy,  the  composition  is 
selected  with  due  regard  to  strength  and  ductibility  as  well  as 
resistance  to  erosion.  Several  manufacturers  have  adopted  steel 
of  high  grade.  In  turbines  patterned  after  the  original  Parsons 
type,  in  which  the  outer  ends  of  the  blades  have  no  support,  a 


COMMERCIAL  ASPECT  OF  THE  TURBINE  351 

steel  lacing  is  employed,  twisted  and  interwoven  between  the 
blades  near  their  outer  ends.  By  these  means,  and  the  removal 
of  causes  leading  to  the  rubbing  of  the  blades  of  the  rotor  against 
the  stationary  parts,  blade  troubles  have  been  almost  entirely 
eliminated,  where  the  machines  have  proper  care  in  the  power 
plant. 

A  more  substantial  blade  support  is  desirable,  however,  and  a 
marked  improvement  has  been  made  in  the  construction  covered 
by  the  Fulleger  and  Sankey  patents,  already  described  in  con- 
nection with  the  Allis-Chalmers  and  Willans  and  Robinson  tur- 
bines. The  shroud  ring  employed  in  this  type  effectually  shields 
the  blades  against  destruction,  even  when  there  is  rubbing  con- 
tact or  a  foreign  substance  enters  the  turbine. 

An  interesting  example  of  the  effectiveness  of  this  is  afforded 
by  a  5,500  Kw.  unit  installed  at  the  Kent  Avenue  power  house  of 
the  Brooklyn  Rapid  Transit  Company,  by  the  Allis-Chalmers 
Company.  When  the  turbine  had  been  running  several  days  after 
it  was  first  started,  the  casing  was  opened  for  inspection.  It  was 
found  that  the  blading  had  been  rubbing  hard  against  the  casing 
for  about  one  third  the  length  of  the  turbine,  but  without  in  any- 
way injuring  the  machine.  Later  the  blading  had  a  still  more 
severe  test.  By  some  means  a  large  jackknife  had  been  left 
inside  the  turbine.  One  of  the  knife  blades  had  gotten  between 
the  spindle  and  the  shroud  of  the  first  row  of  stationary  blades 
and,  acting  like  a  lathe  tool,  had  cut  into  the  body  of  the  drum 
for  a  width  of  about  three  eighths  and  a  depth  of  about  three 
sixteenths  of  an  inch.  This  cut  loosened  the  calking  strip  which 
held  the  ring  of  blades  in  place  and  the  latter,  under  the  influ- 
ence of  centrifugal  force,  had  bent  outward,  so  that  the  channel- 
shaped  shroud  ring  had  rubbed  hard  in  the  bore  of  the  cylinder, 
and  the  flanges  of  the  ring  had  been  worn  down  almost  to  the 
heads  of  the  rivets  which  hold  the  ring  to  the  blades.  Not  a 
single  blade,  however,  had  become  loosened  or  injured. 

In  the  Curtis  Turbine  the  shroud  ring,  and  also  the  base  ring 
of  the  bucket  segments,  are  wider  than  the  blades  themselves, 
thus  effectually  protecting  them.  The  construction  in  this  respect 
is  an  improvement  over  the  earlier  form  used. 


352  STEAM  TURBINES 

Blade  Erosion. 

During  the  lifetime  of  a  reciprocating  engine,  there  is  con- 
tinually increasing  steam  leakage,  because  of  the  wear  of  the 
valves,  piston  rings  and  cylinder.*  At  best  this  loss  is  considerable, 
and  in  order  to  maintain  the  economy  of  the  engine  the  valves 
must  occasionally  be  scraped  to  their  seats,  the  cylinder  rebored, 
and  the  piston  rings  refitted. 

In  the  turbine  there  are  no  corresponding  wearing  parts  and 
practically  the  only  deterioration  that  can  affect  the  steam  con- 
sumption comes  from  the  cutting  action  of  the  steam  or  water 
upon  the  blades.  Experience  thus  far  indicates  that  blade  erosion 
will  not  prove  a  serious  matter ;  but  the  turbine  must  pass  through 
a  longer  trying-out  period  than  it  yet  has  to  demonstrate  whether 
a  drop  in  steam  economy  is  to  be  expected  from  this  cause,  when 
a  machine  has  had  a  long  period  of  service.f 

Erosion  Caused  by  High  Velocity  and  Moisture. — To  test  the 
tendency  of  steam  flowing  at  different  velocities  to  erode  the 
surfaces  of  buckets,  Francis  Hodgkinson  experimented  at  the 
Westinghouse  Machine  Company's  plant  with  hard-drawn  delta 
metal  blades  exposed  to  two  steam  jets.  The  velocity  of  one  jet 
was  about  2,900  feet  per  second  and  of  the  other  about  600  feet 
per  second.  The  blades  were  continuously  exposed  to  the  jets 
for  128  hours.  Those  subjected  to  the  higher  velocity  were 
stripped  and  eroded,  while  those  subjected  to  the  lower  velocity 
were  not  injured. :f 


*In  1904-05  the  Steam  Research  Committee  of  the  Institution  of  Mechanical  En- 
gineers, England,  investigated  the  leakage  of  valves  and  pistons  of  a  small  slide- 
valve  compound  engine.  The  tests  were  carried  out  under  all  manner  of  conditions, 
and  the  results  show  that  with  well-fitted  valves  the  leakage  may  amount  to  over  20  per 
cent  and  is  rarely  less  than  4  per  cent,  depending  upon  the  steam  pressure,  speed,  lap 
of  valves,  etc.  Other  types  of  valves  might  be  either  better  or  worse. 

f  A  500  Kw.  Parsons  turbine  was  installed  at  the  plant  of  the  Cambridge  Electrical 
Supply  Company,  England.  After  it  had  operated  about  a  year,  it  was  tested  by 
Professor  Ewing  and  showed  a  result  of  25  pounds  per  Kw.  hour,  normal  load;  and 
24.4  at  a  slight  overload.  The  factory  test  of  this  machine  showed  a  result  of  24.1 
pounds  per  Kw.  hour.  In  the  later  tests,  however,  besides  running  with  wet  steam, 
the  turbine  was  driving  its  own  air  and  circulating  pumps,  and  the  steam  for  these 
was  charged  to  the  turbine.  In  the  test  at  the  builder's  works  the  turbine  did  not 
drive  its  own  pumps.  There  have  also  been  some  other  tests  of  turbines  after  com- 
paratively short  periods  of  operation,  but  as  yet  no  results  have  been  published  of  tests 
made  after  turbines  have  been  in  longer  operation,  say  for  a  period  of  10  years. 

JPaper  by  Francis  Hodgkinson  before  the  A.  S.  M.   E.  in  1904. 


COMMERCIAL  ASPECT  OF  THE  TURBINE  353 

It  also  appears  to  be  well  established  that  erosion  is  greatly 
increased  by  the  presence  of  moisture  in  steam,  especially  when 
flowing  at  high  velocities,  and  this  may  account  for  the  rapid 
wear  in  the  case  of  the  high-velocity  steam  in  the  Hodgkinson 
experiment.  Such  action  is  corroborated  by  the  manner  in  which 
steam  injectors  invariably  wear.  The  steam  nozzles  of  injectors 
are  seldom  eroded,  although  steam  flows  through  them  at  veloci- 
ties exceeding  1,500  feet  per  second;  but  the  combining  nozzles, 
through  which  the  feed  water  and  condensed  steam  pass  at  a 
more  moderate  rate,  are  often  so  badly  scored  that  they  must  be 
renewed.  Inquiry  of  several  manufacturers  of  injectors  has 
brought  replies  showing  that  but  little  trouble  is  experienced  with 
the  steam  nozzles. 

Erosion  in  De  Laval  Turbines. — In  the  De  Laval  turbine,  in 
which  enormous  steam  velocities  are  realized,  there  is  occasional 
cutting  of  the  blades,  when  the  conditions  of  water  or  steam  are 
not  favorable.  This  was  commented  upon  in  a  paper  before  the 
A.  S.  M.  E.  in  1904,  by  E.  S.  Lea,  then  of  the  De  Laval  Steam 
Turbine  Company,  who  stated  that  there  have  been  a  few  in- 
stances where  buckets  have  worn  out  in  a  year,  necessitating  re- 
placement. In  other  cases  the  wear  has  been  very  slight,  even  in 
a  run  of  four  or  five  "years.  The  wear  affects  only  the  steam  inlet 
side  of  the  buckets  and  hence  does  not  impair  the  efficiency  to  a 
great  extent.  In  tests  upon  a  turbine  of  100  horse-power,  where 
the  edge  of  the  buckets  had  been  worn  away  about  one  sixteenth 
inch,  the  steam  consumption  was  about  five  per  cent  higher  than 
with  new  buckets. 

Erosion  in  Parsons  Turbines. — In  the  Parsons  turbine,  where 
steam  velocities  are  low,  the  trouble  from  erosion  appears  to  be 
almost  entirely  absent,  such  cutting  as  occurs  being  slight  and 
mostly  in  the  low-pressure  end,  where  the  steam  is  moist.  Some 
time  ago  articles  were  published  in  certain  technical  journals,  in 
which  were  illustrations  of  badly  scored  Parsons  buckets,  and 
the  impression  was  conveyed  that  they  were  samples  of  the  condi- 
tion that  turbine  buckets  might  be  expected  to  get  into.  The 
author  succeeded  in  running  down  the  source  of  the  information 
and  found  that  the  blades  illustrated  had  been  taken  from  a  tur- 
bine which  had  become  injured  by  contact  of  the  moving  vanes 


354 


STEAM  TURBINES 


with  the  casing,  and  that  the  cutting  had  undoubtedly  been  done 
by  particles  of  steel  broken  off  and  blown  through  with  the 
steam. 

To  further  investigate  this  important  subject  letters  were  writ- 
ten to  American  engine  builders  and  to  a  number  of  engineers  in 
England,  where  the  turbine  has  been  used  longer  than  in  this 


Fig.   13.     Appearance  of  Turbine  Rotor  after  five  years'   Service. 

country,  asking  for  definite  information  in  regard  to  blade 
erosion.  It  was  expected  that  the  engine  builders,  at  least, 
would  be  well  informed  upon  turbine  difficulties;  but  no  in- 
formation was  secured  from  them,  nor  from  any  other  source,  to 
indicate  that  the  question  of  erosion  need  cause  apprehension. 
Some  erosion  does  occur  when  the  conditions  are  right  for  it, 
even  in  the  Parsons  type  of  turbine,  with  its  low  steam  velocities ; 


COMMERCIAL  ASPECT  OF  THE  TURBINE  355 

but  engineers  in  general  hold  to  the  opinion  that  it  is  not  a  matter 
of  great  moment. 

Experience  with  Westinghouse  Turbines. — The  first  turbine  of 
the  Parsons  type  to  be  put  into  practical  use  in  this  country  was 
installed  at  the  air  brake  works,  Wilmerding,  Pa.,  in  1899.  After 
it  had  been  in  continuous  service  24  hours  a  day  for  over  five 
years  the  blades  were  carefully  inspected  for  wear.  Fig.  13  is  a 
view  of  a  portion  of  the  rotor  at  the  low-pressure  end,  where  the 
steam  is  the  most  moist  and  cutting  would  be  expected,  if  any- 
where. In  Fig.  14  is  a  view  of  a  section  of  the  blading  in  the 
upper  half  of  the'  casing.  Both  views  show  the  blades  to  be  in 
good  condition,  and  the  company  states  that  the  only  wear  that 
could  be  detected  was  in  the  case  of  certain  vanes  which  were 
slightly  out  of  line  with  the  remaining  ones  of  their  particular 
ring.  One  of  these  was  broken  out  and  is  shown  in  Fig.  15, 
photographed  beside  a  new  blade  for  comparison.  Such  wear  as 
occurred  was  chiefly  at  points  a,  a.  The  edges  of  the  blade  were 
worn  to  a  knife  edge.  The  steam  supplied  to  this  turbine  is  said 
to  have  been  excessively  wet,  the  feed  water  extremely  acid,  and 
to  carry  so  much  sediment  that  the  turbine  had  frequently  to  be 
cleaned  out  by  air  blast. 

Wear  of  Curtis  Blades. — The  only  information  the  author  has 
seen  upon  erosion  in  the  Curtis  turbine  has  been  given  by  Chas.  B. 
Burleigh  of  the  General  Electric  Company.*  He  says: 

"The  only  data  we  can  present  are  from  experience,  to  the 
effect  that  after  the  most  exhaustive  tests  under  most  trying 
conditions  no  appreciable  wear  can  be  detected;  and  of  some 
100  turbines  of  an  approximate  aggregate  capacity  of  some 
103,600  Kw.,  all  of  which  are  in  commercial  operation  and  many 
of  which  have  been  in  almost  continuous  commercial  operation  for 
nearly  two  years,  we  have  been  unable  to  detect  the  slightest 
evidence  of  wear. 

"I  had  an  opportunity  last  week  of  examining  the  interior  of  a 
2,000  Kw.  turbine  that  had  been  in  service  for  a  year  on  railroad 
work,  in  company  with  a  prominent  New  England  manufacturer, 
and  we  were  unable  to  detect  the  slightest  evidence  of  any  change 

*  Paper  on  the  Curtis  turbine  before  the  New  England  Railroad  Club,  April,  1905. 


356 


STEAM  TURBINES 


Fig.   14.     Blading  in  upper  half  of  Casing. 


Fig.   15.     Old    Blades    compared    with    new    Blades. 


COMMERCIAL  ASPECT  OF  TPIE  TURBINE  357 

in  the  appearance  of  any  part  of  the  turbine  with  which  the 

steam  came  in  contact Nor  would  this  be  a  matter  of 

serious  moment  if  conditions  were  different,  for  the  reason  that 
if  it  were  necessary  to  replace  every  part  of  a  Curtis  turbine  with 
which  the  steam  comes  in  contact  the  machine  is  so  constructed 
that  this  could  be  accomplished  without  serious  inconvenience  and 
at  an  expense  not  exceeding  10  per  cent  of  the  first  cost." 

Composition  of  Blades. — There  is  no  doubt  that  the  breakage 
and  erosion  of  blades  depends  to  a  considerable  extent  upon  their 
composition,  which,  as  before  stated,  has  received  a  great  deal  of 
attention.  Ordinary  bronzes  containing  tin  have  their  properties 
too  much  affected  at  the  temperatures  of  superheated  steam  to 
make  them  reliable,  one  of  the  reasons  being  that  tin  melts  at  450 
degrees  F.  Brass  also  weakens  at  high  temperatures,  to  a  less 
extent,  but  has  been  extensively  used  for  blades,  the  alloy  varying 
from  72  parts  copper  and  28  parts  zinc,  to  63  copper  and  37  zinc. 
Its  tensile  strength,  annealed,  is  about  45,000  pounds  per  square 
inch ;  but  by  cold  drawing  this  can  be  increased. 

An  alloy  of  about  65,000  pounds  tensile  strength,  which  with- 
stands erosion  well,  consists  of  80  parts  copper  and  20  parts 
nickel ;  and  it  is  only  slightly  affected  by  temperatures  coming 
within  the  range  of  steam  temperatures  in  practice.  Steel  forg- 
ings,  which  are  used  more  or  less  for  blades,  also  retain  their 
properties  satisfactorily  at  the  temperatures  of  steam. 

One  manufacturer  uses  a  nickel  bronze,  which  is  said  to  be  a 
copper-zinc  alloy  containing  a  small  percentage  of  nickel  and  iron, 
the  latter  to  increase  its  strength  and  wearing  qualities.  This 
bronze  is  the  result  of  much  experimenting  and  its  makers  do  not 
care  to  give  the  exact  composition. 


CHAPTER  XVIII 
CARE  AND  MANAGEMENT. 

The  duties  of  the  engineer  of  a  turbine  plant  are  in  most  re- 
spects like  those  of  the  engineer  of  a  plant  equipped  with  recipro- 
cating engines.  There  are,  however,  special  things  to  be  attended 
to  in  order  to  keep  a  turbine  in  good  running  condition. 

First  of  all  it  must  be  remembered  that  the  turbine  is  a  high- 
speed machine  and  that  if  anything  is  to  happen  to  it  it  will 
happen  suddenly  and  almost  without  warning.  A  turbine  that 
has  frequent  inspection  and  regular  care  will  run  day  in  and  day 
out.  But  if  the  oil  circulation  is  allowed  to  fail,  or  the  step- 
bearing  pump  allowed  to  balk,  or  other  vital  part  to  get  out  of 
order  through  lack  of  attention,  a  shutdown  is  the  inevitable 
result.  An  engineer  must  not  deceive  himself  by  thinking  he  can 
coax  a  turbine  along  which  he  has  not  kept  up  in  condition. 
There  is  no  possibility,  for  example,  of  nursing  a  hot  bearing  on  a 
turbine  as  so  often  done  with  a  reciprocating  engine. 

It  is  generally  held  that  turbines  require  less  care  than  re- 
ciprocating engines,  which  is  true  if  by  "care"  is  meant  the 
actual  labor  expended  upon  the  turbine  itself.  But  if  the  high- 
vacuum  condensing  system  be  counted  in,  it  is  a  fair  question  for 
argument  whether  a  "rotary  engineer"  may  not  be  kept  just  as 
busy  as  a  "reciprocating  engineer."*  It  needs  to  be  emphasized 


*Upon  this  point  C.  J.  Davidson,  chief  engineer  of  power  plants,  the  Milwaukee 
Electric  Railway  and  Light  Company,  writes  the  author  as  follows:  "In  our 
company  there  is  no  great  difference  in  the  extent  and  degree  or  quality  of  attend- 
ance (required  by  turbines  and  reciprocating  engines),  notwithstanding  the  popular 
opinion  to  the  contrary.  Our  experience  has  been  with  turbines  of  the  Curtis  type. 
While  it  may  be  possible  to  realize  the  claims  of  some  of  the  advocates  of  turbines 
relative  to  their  ability  to  be  put  quickly  in  service,  it  is  both  our  experience  and 
observation  that  it  requires  some  time  after  starting  for  the  bucket  wheels  to  find 
their  final  running  position,  due  to  expansion,  and  on  this  account  we  exercise  quite 
as  much  care  as  we  would  do  in  warming  up  a  large  engine. 

"The  step-bearing  pump  must  be  kept  in  constant  operation  or  serious  results  will 
follow.  This  means  eternal  vigilance.  Synchronizing  two  alternators  driven  by 
turbines  is  exceptionally  easy ;  but  great  precision  and  consequent  care  on  the  part 
of  the  attendant  is  necessary,  as  a  comparatively  slight  shock  will  unbalance  these 
machines. 

"In  general  I  should  say  that  the  modern  steam  turbine  is  more  refined  me- 
chanically and  consequently  a  more  delicate  piece  of  apparatus  than  the  reciprocating 
steam  engine,  and  to  insure  its  reliability  of  operation  probably  requires  somewhat 
less  labor  but  correspondingly  greater  skill  than  is  necessary  in  case  of  the  engine. >: 


CARE  AND  MANAGEMENT  359 

that  close  watch  must  be  kept  of  the  oiling  system,  vacuum,  step- 
bearing  pump,  gland  water  (the  former  on  the  Curtis  and  the 
latter  on  the  Parsons),  etc.,  while  the  turbine  is  running;  and 
when  the  turbine  is  not  running,  the  auxiliaries  should  be  gone 
over,  glands  packed,  oiling  system  inspected  and  cleaned,  if 
necessary,  and  valves  and  governor  parts  inspected  for  freedom 
of  movement,  and  for  leaks  and  wear,  respectively. 

In  what  follows  a  few  general  directions  will  be  given,  after 
which  the  handling  of  different  turbines  will  be  considered  sepa- 
rately. 

Starting. — The  same  care  must  be  exercised  in  warming  up  a 
turbine  as  a  steam  engine  of  corresponding  power.  Let  steam 
blow  through  slowly  with  the  turbine  standing  idle.  During  this 
interval  start  the  auxiliaries,  first  the  circulating  pump,  then  the 
hot-well  and  dry-air  pumps  and  the  oil  pump.  Finally,  start  the 
turbine  slowly,  later  bringing  it  up  to  speed.  Start  the'  exciter 
set  after  the  turbine  begins  to  rotate.  As  in  the  case  of  a  re- 
ciprocating engine,  gradual  starting  will  avoid  a  sudden  rush  of 
steam  from  the  boiler,  carrying  water  with  it  and  sweeping  along 
the  water  of  condensation  lodged  in  the  piping. 

If  steam  is  drawn  from  a  superheater,  take  particular  care  in 
handling  the  admission  valve,  for  during  the  warming  up  the 
small  quantity  of  steam  flowing  will  become  considerably  cooled, 
while  the  larger  volume  flowing  when  the  throttle  is  opened  wider 
will  retain  its  superheat  and  the  turbine  will  be  suddenly  exposed 
to  a  temperature  much  higher  than  that  of  the  warming-up 
period.  In  warming  up  with  superheated  steam,  rotate  the 
spindle  slowly  for  a  long  enough  period  to  ensure  that  all  parts 
are  brought  up  uniformly  to  their  normal  temperature  before 
bringing  up  to  speed. 

The  chief  concern  of  the  engineer  in  starting  should  be  to  do 
nothing  to  produce  sudden  changes  of  temperature  in  the  tur- 
bine. Where  turbines  are  standing  idle  part  of  the  time,  but  are 
liable  to  be  called  into  service,  it  is  a  good  plan  to  keep  them  warm 
all  the  time  by  allowing  a  small  quantity  of  steam  to  enter  con- 
tinuously, preferably  through  a  by-pass  in  the  throttle.  In  this 
way  a  much  quicker  start  can  be  made,  without  danger  of  a  mis- 
hap, or  of  the  rotor  being  out  of  balance  and  vibrating.  A  rapid 


360  STEAM  TURBINES 

start  can  be  made  more  easily  with  Curtis  than  with  Parsons  tur- 
bines and  if  a  Curtis  turbine  has  been  kept  warm  it  can  be 
brought  up  to  speed  in  two  or  three  minutes  in  an  emergency. 
In  general,  however,  10  to  15  minutes  should  be  taken  in  warm- 
ing up  and  starting  either  a  Curtis  or  Parsons  turbine,  and  if 
the  auxiliaries  are  put  into  operation  at  the  same  time  it  will 
usually  require  about  the  same  interval  to  get  them  running 
regularly. 

Shutting  Down. — Partly  close  the  throttle  before  reducing  the 
load  on  the  generator,  so  the  turbine  can  be  brought  under  instant 
control  in  case  it  should  speed  up  when  the  load  is  thrown  off. 
This  cannot  happen,  of  course,  if  the  safety  stop  is  operative. 
After  closing  the  throttle  it  is  well  to  trip  the  stop  motion  to  test 
its  action.  After  shutting  off  steam  close  the  condenser  valve ;  or 
if  the  turbine  is  connected  to  an  independent  condenser,  stop  the 
air  pump,  hot-well  pump  and  circulating  pump.  It  is  not  un- 
common for  a  turbine  rotor,  running  in  vacuum  and  with  no  load, 
to  continue  to  rotate  for  from  30  to  60  minutes  after  steam  is 
shut  off.  The  speed  can  be  checked  by  opening  the  drains, 
admitting  air  to  the  casing,  and  by  leaving  the  current  on  the 
generator  fields.  If  the  turbine  has  an  independent  load,  in- 
stead of  running  in  parallel  with  others,  this  can  be  used  to 
quickly  check  the  speed. 

Condensing  Apparatus. — The  turbine  engineer,  who  has  a  high- 
vacuum  surface  condenser  and  connected  apparatus  under  his 
care,  will  find  the  turbine  itself  to  be  the  least  source  of  his 
troubles.  A  loss  of  an  inch  or  two  in  vacuum  in  a  reciprocating 
engine  plant,  where  26  inches  is  considered  a  good  vacuum,  is 
not  a  serious  matter.  But  in  a  turbine  plant,  where  27  or  28 
inches  or  more  are  carried,  a  drop  of  an  inch  or  two  in  vacuum 
means  a  large  increase  in  steam  consumption,  as  explained  in 
Chapter  XIX.  It  is  therefore  a  much  more  important  matter  to 
keep  the  condensing  system  of  a  turbine  plant  up  to  a  high  state 
of  efficiency  than  in  the  case  of  an  engine  plant,  and  it  is  also 
a  much  more  difficult  matter  to  do  so,  because  of  the  greater 
chance  for  air  leakage  through  glands,  joints  and  relief  valve. 
It  is  a  temptation  to  let  air  leaks  go  and  cover  up  their  existence 
by  pushing  the  air  and  circulating  pumps,  with  consequent  addi- 


CARE  AND  MANAGEMENT  361 

tion  to  operating  expenses.  But  the  painstaking  engineer  will 
not  be  satisfied  to  do  business  in  this  way. 

More  or  less  trouble  is  experienced  from  the  carbonization  of 
the  oil  in  the  ports  and  valve  chambers  of  the  dry-air  pump. 
One  plan  for  overcoming  this  is  to  provide  an  additional  oil  cup 
of  the  positive-feed  type  for  the  air  cylinder,  and  to  use  this  to 
feed  in  soap  suds  along  with  the  oil.  The  trouble  can  also  be  re- 
duced by  having  the  jacket  cooling  water  as  cold  as  possible  and 
forcing  a  large  quantity  through  the  jacket.  Mineral  oil  should 
be  used  of  the  grade  designed  for  gas  engines  and  air  compressors. 

Changing  from  Condensing  to  Non-condensing. — When  run- 
ning condensing,  the  exhaust  end  and  exhaust  passages  of  the 
turbine  are  cool,  but  if  a  change  is  made  to  non-condensing  the 
temperature  of  the  steam  in  these  passages  will  rise  at  once  to 
above  212  degrees  and  the  quantity  of  steam  flowing  will  also 
increase.  If  the  change  is  made  suddenly,  the  turbine  will  be 
subjected  to  wide  temperature  changes  and  care  must  be  exer- 
cised to  shut  off  the  condenser  as  gradually  as  possible  to  avoid 
this.  When  changing  from  non-condensing  to  condensing,  the 
weight  of  steam  flowing  diminishes  and  the  cooling  effect  will 
not  be  as  marked  as  was  the  heating  effect  in  the  other  case. 

Operating  the  De  Laval  Turbine.* 

Starting. — Upon  first  starting,  after  erecting  or  after  a  long 
shutdown,  the  bearings  should  be  flooded  with  oil,  the  amount 
being  gradually  reduced  to  the  normal  quantity.  The  oil  reser- 
voirs on  the  self-oiling  bearings  should  be  filled  until  the  oil 
stands  between  the  red  marks  on  the  gauge  glass.  The  small 
oil  valves  on  the  governor  valve  should  be  filled  with  cylinder 
oil,  the  valve  stems  then  pressed  down,  thus  allowing  the  oil  to 
pass  into  the  governor  valve.  Steam  is  then  turned  on,  and  the 
governor  valve  and  wheel  case  allowed  to  become  thoroughly 
heated.-  Before  doing  this,  however,  the  nozzle  valves  should  be 
opened  about  a  half  turn;  otherwise  they  will  stick  when  the 
wheel  case  becomes  hot.  The  turbine  should  be  started  gradually 
so  as  to  give  the  bearings  time  to  heat  thoroughly.  More  time 

*Abridged  from  directions  furnished  by  the  De  Laval  Steam  Turbine  Company. 


362  STEAM  TURBINES 

is  required  for  this  in  the  larger  turbines  than  in  the  smaller.  As 
soon  as  the  turbine  starts,  the  self-oiling  bearings  must  be  ex- 
amined to  see  if  the  oil  rings  run  properly.  If  the  turbine  is 
running  condensing,  the  condenser  should  be  started  first.  If 
starting  with  no  load  it.  is  well  to  start  with  a  low  vacuum,  say 
from  24  to  25  inches.  As  soon  as  the  load  is  put  on,  the 
vacuum  should  be  raised  to  its  maximum. 

Shutting  Dozvn. — When  a  machine  running  non-condensing 
is  to  be  stopped,  the  throttle  valve  should  be  closed  and  the 
lubricator  shut  off  as  soon  as  the  machine  has  come  to  a  stand- 
still. If  the  turbine  is  running  condensing,  and  if  operating  the 
water  and  air  pumps,  either  directly  or  indirectly,  the  air  cock 
on  the  exhaust  end  of  the  turbine  wheel  case  should  be  opened 
before  the  throttle  valve  is  shut  off. 

General  Care  of  Turbine. — The  usual  precautions,  with  which 
engineers  are  familiar,  should  be  taken  to  keep  the  oiling  ar- 
rangements in  working  order.  The  sight-feed  lubricator  must 
be  kept  clean  and  the  oil  in  the  self-oiling  bearings,  and  accumu- 
lating in  the  gear  case,  drawn  off  and  filtered  as  often  as  neces- 
sary. Particular  attention  should  be  given  to  the  oiling  of  the 
governor  mechanism,  and  especially  the  contact  surfaces  between 
the  governor  pin  and  the  plunger  on  the  bellcrank.  The  high- 
speed bearings  should  be  removed  and  examined  at  intervals. 
Should  a  bearing  run  hot,  it  should  be  taken  out,  the  oil-grooves 
cleaned  and  if  any  bright  or  black  spots  appear,  they  should  be 
removed  with  a  scraper. 

The  strainer  above  the  governor  valve,  to  prevent  foreign 
particles  from  entering  the  turbine,  should  be  removed  and  ex- 
amined at  least  once  a  month. 

If  the  turbine  speed  is  too  high,  the  brass  nut  holding  the 
governor  springs  should  be  screwed  out,  or  if  the  speed  is  too 
low,  the  nut  should  be  tightened.  It  is  well,  every  time  a  tur- 
bine is  started,  to  press  down  the  bellcrank,  to  ascertain  that 
these  parts  do  not  stick;  and  when  fully  depressed,  the  governor 
valve  should  shut  off  steam  entirely,  or  at  least  within  a  few 
pounds. 

To  keep  the  gears  in  proper  condition,  the  teeth  should  be 
cleaned  occasionally  when  the  machine  is  not  running.  Kero- 


CARE  AND  MANAGEMENT  363 

sene  and  a  metal  brush  are  the  best  for  this  purpose.  The  gear 
case  should  also  be  cleaned  at  the  same  time  and  the  gears  well 
lubricated. 

If,  for  any  reason,  the  gears  have  to  be  taken  out  of  the  case, 
the  engineer  should  secure  special  directions  from  the  manu- 
facturers, relating  to  the  adjustment  of  the  gears  as  well  as  to 
their  removal.  The  gears  need  to  be  kept  in  perfect  adjustment, 
as  their  life  depends  to  a  considerable  extent  upon  this  being  done. 

Operating  the  Parsons  Turbine. 

Starting  and  Stopping. — It  is  the  rule  of  many  engineers  never 
to  allow  a  condensing  engine  to  run  without  a  vacuum,  either  in 
starting  or  shutting  down,  if  it  can  be  avoided.  They  follow  the 
same  rule  in  operating  turbines  as  in  the  case  of  engines  and 
start  the  condenser  pumps  while  the  turbine  is  warming  up  and 
then  bring  the  turbine  up  to  speed  with  the  vacuum  on.  If 
the  turbine  is  a  Westinghouse  an  objection  to  this  method  lies 
in  the  fact  that  the  glands  of  this  turbine,  which  are  water- 
packed,  do  not  become  sealed  until  the  machine  is  in  operation ; 
and  if  the  turbine  is  started  when  subject  to  vacuum  there  will 
be  leakage  of  cold  air  in  through  the  glands,  tending  to  set  up 
unequal  temperature  conditions  in  the  turbine. 

The  author  has  not  learned  of  any  trouble  from  this  cause,  but 
if  it  is  desired  to  avoid  this  condition  the  order  of  starting  is  to 
start  the  circulating  pump  while  the  turbine  is  warming  up ; 
then  gradually  start  the  turbine  and  exciter  set.  When  the  speed 
increases  to  the  point  where  the  load  begins  to  come  on,  turn  on 
the  gland  water,  start  the  hot-well  pump  and  dry-air  pump,  and 
bring  the  turbine  and  exciter  up  to  speed  together.  In  shutting 
down  turn  off  the  water  from  the  glands  as  soon  as  the  vacuum 
drops,  and  open  the  drip  pipes. 

At  the  plants  of  the  Hartford  Electric  Light  Company,  Hart- 
ford, Conn.,  where  they  have  had  the  longest  experience  with 
Westinghouse  turbines  of  any  company  in  the  country,  it  is  the 
practice  to  start  up  non-condensing  until  the  glands  have  been 
closed,  or  nearly  so,  before  throwing  in  condensing. 

Directions  by  an  Engineer. — W.  H.  Damon,  Springfield,  Mass., 


364  STEAM  TURBINES 

engineer  of  the  United  Lighting  Company  of  that  city,  has 
written  the  author  regarding  the  care  of  Westinghouse-Parsons 
turbines  as  follows : 

We  have  three  1,000  Kw.  Westinghouse-Parsons  turbines  connected 
with  jet  condenser.  They  have  given  us  no  trouble  at  all,  sometimes 
running  from  Sunday  to  Sunday  without  stopping. 

In  starting  up  we  warm  up  the  turbine,  start  the  dry-air  pump,  and 
then  the  injection  pump.  The  turbine  is  then  started  slowly,  followed 
by  the  exciter,  which  is  steam-driven.  The  turbine  and  exciter  are 
brought  up  to  speed  at  about  the  same  time,  taking  from  10  to  15  min- 
utes to  get  up  to  speed.  If  you  start  too  fast  there  is  too  great  a 
vibration. 

After  the  turbine  is  up  to  speed  and  the  load  on;  all  the  attention  it 
needs  is  to  watch  the  oil  supply  on  the  bearings  and  the  gland  water. 

The  auxiliaries  need  more  care  than  the  turbine.  We  have  had  some 
trouble  from  the  oil  carbonizing  in  the  valve  chambers  of  the  dry-air 
pump,  stopping  up  the  ports  and  causing  the  valves  to  stick,  but  have 
overcome  this  to  a  great  extent.  Otherwise  the  care  of  the  outfit  is  sim- 
ply keeping  things  clean,  cleaning  the  oil  strainers  on  the  turbine,  keeping 
the  governor  from  getting  gummed  and  the  pilot  valve  in  good  condition. 
The  oiling  system  should  be  given  close  attention,  being  careful  not  to 
pump  any  air  into  the  system  and  having  plenty  of  oil  in  the  suction  tank. 

In  stopping  close  the  throttle  and  shut  down  the  exciter  and  con- 
densers, being  sure  to  shut  off  the  gland  water  as  the  water  might  other- 
wise get  into  the  oil  as  the  vacuum  falls. 

We  use  an  auxiliary  oil  pump  to  ensure  a  good  supply  of  oil  on  the 
bearings  when  the  turbine  is  running  at  a  slow  speed. 

Care  of  the  Turbine. — In  the  Parsons  turbine  the  spindle 
bearings  support  the  weight  of  the  drum  and  this  weight,  in 
connection  with  the  high  speed  of  the  journals,  causes  the  bear- 
ings to  run  so  hot  that  the  hand  can  scarcely  be  held  on  them. 
Their  high  temperature  makes  necessary  the  cooling  coil  for 
the  oil  and  this  must  be  cleaned  as  often  as  required  to  keep  the 
coil  surfaces  effective.  This  alternate  heating  and  cooling  of 
the  oil  makes  some  oils,  which  otherwise  would  be  good  lubri- 
cants, poorly  adapted  for  turbine  work.  The  heating  tends  to 
decompose  them  and  if  they  contain  paraffine  this  will  be  de- 
posited on  the  surfaces  of  the  coil  and  in  the  oil  passages  dur- 
ing the  cooling  process.  Some  oils,  also,  have  a  tendency  to 
form  an  emulsion  when  they  become  mixed  with  water,  which 


CARE  AND  MANAGEMENT  365 

might  happen  in  case  of  leakage  from  the  turbine  glands.     This 
emulsion  is  like  jelly  and  chokes  up  the  coil. 

In  the  general  care  of  the  turbine,  the  governor  parts  and 
connections  with  the  primary  and  secondary  admission  valves 
must  be  regularly  inspected  to  see  that  they  do  not  become 
gummed,  and  it  must  be  seen  that  the  pilot  valves  work  freely 
and  are  in  good  condition.  Occasional  inspection  of  the  blading 
is  advisable,  at  which  times  the  blade  channels  may  be  cleaned, 
if  required.  When  the  machine  is  running,  give  the  oiling  system 
close  attention,  making  sure  there  is  enough  oil  in  the  suction 
tank  to  avoid  air  being  drawn  into  the  system.  Try  the  pet  cock 
on  each  bearing  frequently  to  see  that  the  oil  is  circulating 
properly.  In  the  base  of  the  turbine  is  an  oil  strainer  which 
should  be  removed  and  cleaned  every  few  days  and  which  can  be 
done  while  the  turbine  is  in  operation.  The  oiling  system  and 
the  water  supply  to  the  glands  and  cooling  coil  are  the  three 
things  that  require  regular  attention  when  the  turbine  is  running. 

Operating  the  Curtis  Turbine. 

Practically  the  only  feature  of  a  Curtis  Turbine  (aside  from 
the  condensing  apparatus)  which  requires  care  or  attention  dif- 
ferent from  and  in  addition  to  that  which  would  be  given  a 
steam  engine  is  the  high-pressure  hydraulic  system  for  the  step 
bearing.  Double-acting  duplex  pumps  are  used,  usually  in  con- 
nection with  an  accumulator,  and  so  much  depends  upon  the 
maintenance  of  pressure  in  the  step  bearing,  and  the  pump  is 
working  under  such  high  pressure,  that  unusual  care  must  be 
taken  to  keep  the  pumps  packed  and  in  good  order  and  to  see 
that  they  are  regularly  inspected  when  in  operation. 

Directions  for  Care  of  Turbine. — The  author  has  talked  and 
corresponded  with  many  turbine  engineers  in  regard  to  the  man- 
agement of  their  plants,  and  among  the  letters  received  is  one 
from  A.  A.  Leavitt,  engineer  of  the  Gloucester,  Mass.,  Electric 
Company,  in  which  are  the  following  concise  directions  for 
handling  Curtis  turbines: 

1.  The  air  and  circulating  pumps  and  all  piping  connections  to  the  same 
must  receive  frequent  attention,  as  it  is  of  the  greatest  importance  that  the 
vacuum  carried  be  as  high  as  possible. 


366  STEAM  TURBINES 

2.  The  pressure  pumps  must  be  kept  in  first-class  condition  and  piping 
examined  frequently  to  ensure  its  being  in  good  condition.    As  the  pressure 
carried  is  high,  usually  about  500  pounds  per  square  inch  between  pumps 
and  accumulator,  and  about  200  pounds  at  the  step  bearing,  and  as  this 
step  bearing  is  what  carries  the  whole  machine,  the  importance  of  attention 
to  this  feature  is  evident.    It  requires  great  care  in  packing  the  step-bearing 
pumps  to  ensure  a  steady,  uniform  full  stroke. 

3.  The  oiling  system  must  be  kept  tight,  as  a  small  leak  will  not  only 
make  a  machine  look  unsightly,  but  will  materially  affect  the  operation  of 
it  by  the  oil  dropping  down  onto  the  collector  rings  and  causing  sparking. 

4.  The  brushes  and  collector  rings  must  be  kept  absolutely  clean  and 
perfectly  adjusted,  to  ensure  steady  voltage;  for  after  they  start  to  spark 
the  voltage  will  be  very  unsteady. 

5.  The  governor  must  be  kept  in  the  best  possible  condition  to.  ensure 
steady  speed.     As  the  governors  of  these  machines  all  run  at  high  speed 
the  parts  wear  quite  rapidly  and  this  wear  should  be  detected  and  remedied 
by  making  the  necessary  adjustments. 

6.  In  starting,  the  turbine  should  be  given  time  to  warm  up  and  the 
parts  expand  to  working  conditions,  especially  if  a  high  degree  of  super- 
heat is  used.    The  step-bearing  pumps  are  first  started,  to  give  the  accumu- 
lator time  to  rise  to  its  position,  and  then  the  circulating  pump  is  started, 
then  the  vacuum  pump  and  lastly  the  oil  pumps.    After  these  are  all  work- 
ing properly  the  turbine  is  started.    The  exciter  set  is  started  and  the  cur- 
rent put  on  the  turbine  fields  when  the  turbine  is  up  to  about  half  speed. 

7.  In  shutting  down,  shut  steam  off  the  turbine  first,  then  stop  the  air 
pump,  the  circulating  pump,  and  last  the  oil  pumps  and  step-bearing  pumps. 
When  the  turbine  has  come  to  rest,  it  should  be  carefully  gone  over  and 
scrupulously  cleaned,  the  same  as  any  dynamo,  as  there  is  nothing  which 
will   collect    dirt   any    faster   than    electric   machinery   and    there    is    no 
machinery  to  which  dirt  is  any  greater  detriment. 

Practice  at  a  Large  Turbine  Station. — At  the  L-Street  station 
of  the  Edison  Lighting  Company,  South  Boston,  Mass.,  are 
four  5,000  Kw.  Curtis  turbines,  each  with  condenser  in  its  base 
and  auxiliary  apparatus  ranged  about  the  turbine  on  the  same 
floor  level  as  the  turbine.  In  Fig.  5,  page  379,  is  a  view  of  one 
of  these  units,  with  its  group  of  auxiliaries.  Cooling  water  is 
supplied  to  the  condenser  by  a  steam-driven  centrifugal  pump. 
The  wet  or  hot-well  pump  is  an  electrically  driven  centrifugal 
pump  placed  in  a  pit  below  the  floor  level.  The  air  pump  has  a 
single  cylinder,  steam  driven.  There  is  a  boiler  feed  pump  for 
each  turbine  unit,  which  discharges  into  a  heater  where  the  feed 


CARE  AND  MANAGEMENT  367 

water  is  heated  by  steam  from  the  auxiliaries  and  from  an  inter- 
mediate stage  of  the  turbine,  from  which  it  is  taken  direct. 

The  author  is  enabled  to  publish  a  few  notes  on  the  practice 
followed  in  the  operation  of  the  several  units  at  this  station. 

Starting  and  Stopping. — The  order  followed  in  starting  is  first 
to  prime  and  start  the  circulating  pump  so  there  shall  be  no 
possibility  of  overheating  the  condenser  by  admitting  steam  before 
the  cooling  water  is  flowing  through.  The  step  bearing  and  oil 
pumps  are  then  started;  then  the  wet  and  dry  vacuum  pumps; 
and  finally  the  turbine  and  exciter  and  the  feed  pump.  In 
stopping,  the  load  is  taken  off  when  the  speed  of  the  turbine  has 
been  reduced  to  the  point  where  the  voltage  and  cycles  drop.  It 
is  general  practice  to  shut  down  with  the  automatic  valve  instead 
of  with  the  throttle,  to  test  its  reliability.  The  pumps  are  shut 
off  in  the  opposite  order  from  which  they  are  started,  but  the 
step  bearing  pump  must  of  necessity  be  kept  in  operation  until 
the  turbine  ceases  to  rotate,  which  requires  about  J4  hour  at  this 
station. 

Condenser  Pumps. — In  order  to  prime  the  circulating  pump  i 
small  air  pump  or  air  ejector  must  be  used,  the  latter  type  being 
employed  at  this  station. 

The  wet  pump  is  always  started  before  the  dry  air  pump,  to 
remove  any  water  that  may  have  collected  in  the  condenser  during 
the  shut-down  period,  such  as  might  occur  if  there  were  a  leaky 
throttle  or  if  the  step-bearing  water  discharged  into  the  condenser, 
as  is  the  case  in  this  station.  When  the  wet  pump  has  reduced 
the  water  to  a  safe  level  the  dry  pump  is  started.  By  following 
this  order  the  danger  of  injury  to  the  dry  air  pump,  by  drawing 
a  large  volume  of  water  into  the  air  cylinder,  is  avoided. 

The  wet  pumps  are  of  the  centrifugal  type  and  in  their  opera- 
tion it  is  found  important  to  keep  the  glands  of  the  impeller 
shaft  well  packed  to  prevent  air  leaks.  Such  leaks,  if  considerable 
in  amount,  reduce  the  effectiveness  of  the  pumps  and  cause  water 
to  accumulate  in  the  condenser.  Care  is  also  taken  that  these 
pumps  do  not  run  without  a  partial  supply  of  water  to  act  as  a 
lubricant  for  the  rubbing  parts. 

Oiling  System. — The  gravity  oiling  system  is  employed  for  all 


368  STEAM  TURBINES 

the  main  bearings,  except  the  step  bearing.  Oil  is  pumped  from 
a  receiving  tank  to  an  elevated  supply  tank,  from  which  it  flows 
to  a  small  distributing  reservoir  at  the  top  of  the  turbine.  The 
system  requires  no  special  attention,  therefore,  different  from  that 
required  in  engine  work. 

The  lubricant  used  for  the  step  bearing  is  water.  The  step- 
bearing  pumps  operate  against  a  pressure  of  1,200  pounds,  which 
is  reduced  to  800  pounds  at  the  step  bearing  by  passing  through 
a  pressure  reducer  of  the  baffle  type.  An  accumulator  is  used 
with  each  turbine  to  maintain  a  steady  pressure  and  to  act  as  a 
pressure  storage  in  case  the  pump  fails,  the  capacity  being  suffi- 
cient to  hold  the  pressure  for  10  minutes.  If  a  quick  start  is 
likely  to  be  required,  the  accumulator  is  shut  off  from  the  system 
when  the  turbine  is  stopped,  so  that  the  step-bearing  pressure  will 
be  ready  at  a  moment's  notice. 

No  trouble  has  been  experienced  with  the  step  bearings  in  this 
station.  If  the  water  were  gritty,  however,  the  bearings  would 
wear  down  and  would  need  occasional  adjustment  to  bring  the 
turbine  rotor  into  proper  position,  as  determined  by  clearance 
indicators  on  each  stage.  Much  of  the  foreign  matter  in  the 
water  is  removed  by  the  baffle  above  mentioned,  which  should  be 
occasionally  cleaned.  It  is  found  that  even  with  the  large  accumu- 
lators used  at  this  station  it  is  possible  to  pack  them  so  tight  that 
the  plunger  will  not  drop,  and  it  is  the  practice  to  test  each 
accumulator  daily  for  freedom  of  movement  by  causing  the  ram 
.to  move  through  its  whole  range  of  travel. 

Warming  up  and  Synchronising. — Each  of  these  machines  is 
warmed  up  by  a  special  by-pass  and  three  admission  valves  which 
are  electrically  and  separately  controlled  and  furnish  the  neces- 
sary amount  of  steam  for  warming  up  and  also  starting  the  tur- 
bine and  bringing  it  up  to  speed,  taking,  in  this  case,  only  from 
two  to  five  minutes.  The  exciter  set  is  started  and  the  field  given 
excitation  during  the  early  period  of  raising  the  speed.  In  syn- 
chronizing great  care  is  taken  to  have  the  machine  come  into 
phase  while  its  speed  is  accelerating  instead  of  falling  off.  It  is 
common  experience  that  turbines  which  were  once  in  good  align- 
ment and  balance  may  be  thrown  out  of  balance  by  lack  of  care 
on  the  part  of  the  operator  in  synchronizing.  But  by  following 


CARE  AND  MANAGEMENT  369 

the  method  advocated  above,  severe  shocks  will  be  avoided  and  the 
balance  and  adjustment  of  the  parts  preserved. 

Notes  of  Experience. — C.  E.  Stanton,  chief  engineer  of  the 
Union  Electric  Company,  Dubuque,  la.,  gives  the  results  of  his 
operative  experience  with  Curtis  four-stage  500  Kw.  turbines.* 
The  chief  difficulties  have  been  in  connection  with  the  water 
supply  for  the  step  bearings,  the  gravity  oil  supply  for  lubrica- 
tion, and  the  occasional  sticking  of  the  nozzle  valves.  The  diffi- 
culty in  lubrication  arose  through  an  air  lock  formed  in  the 
gravity  oil  tank,  allowing  air  to  come  into  the  oil  feeder  pipe 
line  and  interfere  with  the  flow  of  oil,  and  was  remedied  by 
venting  the  top  of  the  tank. 

All  water  for  the  step  bearings  in  his  plant  passes  through  a 
strainer  after  leaving  the  pumps,  to  remove  particles  that  might 
clog  up  the  passages  of  the  step  bearings  or  injure  the  latter. 
The  pumps  for  the  service  have  fibrous  packing  and  if  this  is 
left  until  it  loses  its  elasticity  and  becomes  soft,  small  particles 
find  their  way  into  the  strainer  and  soon  choke  the  supply  of 
water  to  the  step  bearings.  Dirt  or  particles  of  packing,  when 
once  in  the  system,  may  find  their  way  into  the  strainers,  even 
after  many  days  or  weeks,  and  the  strainers  must  therefore  be 
cleaned  at  intervals  of  twenty-four  hours. 

The  hydraulic  accumulator  for  the  step-bearing  system,  if 
allowed  to  remain  in  one  position  for  a  considerable  period  of 
time,  was  found  to  rust  fast  and  not  drop,  even  if  all  the  pressure 
was  removed  from  the  system,  thus  defeating  the  object  for  which 
the  accumulator  is  intended.  It  therefore  must  be  tested  fre- 
quently by  allowing  the  ram  to  drop  slowly  and  then  return  to 
its  former  position,  a  test  that  was  made  each  day.  One  other 
precaution  that  should  be  taken  with  accumulators  for  this  work 
is  to  have  some  kind  of  signal,  usually  a  steam  whistle,  which 
will  blow  if  the  accumulator  starts  to  come  down,  thus  notifying 
the  engineer  of  the  failure  of  the  oil  supply. 

In  the  500  Kw.  turbines  there  are  eight  main  nozzle  valves, 
each  with  its  individual  pilot  valve,  which  is  electrically  con- 
trolled. On  any  load  within  the  rated  capacity  of  the  turbine, 
running  condensing,  five  valves  are  all  that  open,  leaving  three 

*Paper  presented  at  meeting  of  Iowa  Electrical  Association,  April,  1906. 


370  STEAM  TURBINES 

valves  which  might  not  open  for  days  at  a  time.  If  these  are 
left  long,  they  will  corrode  and  stick  and  if  a  heavy  overload 
should  come  might  not  open  at  all — or  if  they  did  open  they 
might  remain  in  this  position.  To  obviate  these  troubles  all 
valves  are  opened  and  closed  several  times  each  day  when  start- 
ing the  turbines.  Some  difficulty  was  experienced  in  securing 
suitable  packing  for  the  main  nozzle  valves  which  would  stand  a 
high  degree  of  superheat.  Metallic  packing  was  not  successful 
and  asbestos  ring  packing  is  now  employed,  which  is  satis- 
factory, except  that  the  valve  stems  must  be  repacked  more  fre- 
quently than  would  be  the  case  if  metallic  packing  could  be  used. 


CHAPTER  XIX 
CONDENSING  APPARATUS  FOR  HIGH  VACUUM. 

One  of  the  advantages  of  the  turbine,  from  a  thermodynamic 
standpoint,  is  its  ability  to  utilize  a  high  vacuum  and  expand  steam 
down  to  the  lowest  pressure  that  can  be  attained  in  a  condenser. 
The  only  restriction  to  the  number  of  times  that  steam  may  be  ex- 
panded in  a  turbine  is  that  the  passages  must  be  large  enough 
to  accommodate  the  increased  volume  of  the  steam  at  the  low  pres- 
sures. In  order  to  maintain  a  high  vacuum,  however,  a  large,  ex- 
pensive and  somewhat  elaborate  condensing  system  must  be 
adopted,  which  requires  constant  attention  to  keep  in  a  high  state 
of  efficiency.  On  this  account  the  high-vacuum  feature  of  turbine 
plants  has  not  worked  altogether  to  their  advantage,  and  condens- 
ing outfits  have  generally  given  more  trouble  and  have  been  a 
greater  source  of  expense  than  the  turbine  itself. 

Effect  of  High  Vacuum  with  Steam  Engines. — While  it  is,  on 
the  whole,  thought  desirable  to  provide  a  condensing  system  for 
steam  turbines  for  a  vacuum  of  about  28  inches,  the  gain  in 
a  steam  engine  from  increasing  the  vacuum  above  26  inches  is  so 
slight  as  not  usually  to  warrant  the  extra  expense.  A  compound 
engine,  with  the  usual  cylinder  ratio  of  4  to  1,  will  expand  the 
steam  from  10  to  15  times.  The  volume  of  one  pound  of  steam  at 
150  pounds  pressure,  absolute,  is  3  cubic  feet.  In  expanding  15 
times,  or  to  a  volume  of  45  cubic  feet,  the  terminal  pressure  in  the 
low-pressure  cylinder  would  be  between  8  and  9  pounds,  assum- 
ing cylinder  condensation  to  be  balanced  by  reevaporation.  When 
the  exhaust  valve  opened,  therefore,  the  pressure  would  drop  sud- 
denly to  that  of  the  condenser,  and  the  only  effect  of  a  vacuum 
higher  than  that  represented  by  the  8  or  9  pounds  pressure  in  the 
cylinder  would  be  to  reduce  the  back  pressure  against  the  piston 
during  the  return  stroke. 

If  the  attempt  were  made  to  carry  expansion  too  far  in  a  steam 
engine,  the  low-pressure  cylinder,  valves  and  passages  would  have 
to  be  abnormally  large  and  would  offer  a  great  deal  of  frictional 
resistance.  Under  such  conditions  a  point  would  be  reached  where 


372  STEAM  TURBINES 

the  pressure  of  the  steam  would  not  be  sufficient  to  overcome  the 
frictional  resistances,  to  say  nothing  of  doing  useful  work,  and  the 
expansion  of  the  steam  beyond  this  point  would  therefore  be  a 
dead  loss.  The  increased  condensation  in  the  low-pressure 
cylinder  would  also  be  a  serious  factor. 

Below  are  given  the  volume  of  one  pound  of  steam  correspond- 
ing to  different  "vacuum"  pressures,  indicating  how  impossible  it 
is  to  utilize  these  low  pressures  in  the  steam  engine.  To  expand 

Absolute  Vacuum,  Specific 

Pressure.  Inches.  Volume. 

y2  29  636 

1  28  335 

2  26  174 

3  24  118 

4  22  90 

steam  from  150  pounds  to  1  pound  absolute,  or  to  28  inches 
vacuum,  would  mean  that  the  volume  must  increase  111  times. 
To  carry  the  expansion  to  this  point  in  a  compound  engine,  the 
ratio  of  the  cylinders  would  have  to  be  about  33  to  1 ;  that  is,  the 
diameter  of  the  low-pressure  cylinder  would  be  10^  times  that  of 
the  high-pressure  cylinder — quite  an  impracticable  figure. 

In  the  case  of  the  turbine,  however,  the  steam  may  easily  be  ex- 
panded from  100  to  150  times  without  encountering  any  con- 
structive difficulties. 

Why  a  Turbine  Derives  more  Benefit  from  High  Vacuum  than 
an  Engine. — Fig.  1  illustrates  the  expansion  of  one  pound  of  steam 
from  an  initial  pressure  of  100  pounds  to  the  pressures  indicated, 
and  illustrates  the  difference  between  the  way  in  which  an  engine 
and  a  turbine  benefit  from  a  high  vacuum.  It  shows  the  work 
done  both  before  and  during  expansion,  as  in  an  indicator  dia- 
gram. The  section  of  the  diagram  marked  a-b-c-d-e  represents 
that  part  of  the  energy  of  the  steam  that  might  be  converted  into 
work  by  a  condensing  engine  operating  against  a  back  pressure  of 
four  pounds,  or  a  vacuum  of  about  22  inches.  At  point  c  expan- 
sion has  been  carried  as  far  as  the  size  of  the  engine  cylinder 
permits  and  hence,  when  the  exhaust  valve  opens,  the  pressure 
drops  from  point  c  to  point  d. 


CONDENSING  APPARATUS 


373 


Now  assume  the  back  pressure  to  be  reduced  to  two  pounds,  and 
it  is  evident  that  the  gain  in  power  for  the  engine  would  be  due 
simply  to  the  reduction  in  back  pressure  represented  by  the 


a  b 


GO  lb.  Absolute 


4-pound  back-pressure  line 
2-pouud  back-pressure  line 
Line  of  zero  pressure 
Atmospheric  Line 


— X — 
Fig.  1.     Diagram  Showing  How  the  Turbine  Takes  Advantage  of  High  Vacuum. 

shaded  portion  having  the  length  x  on  the  diagram.  This,  it  will 
be  noticed,  is  but  a  small  percentage  of  the  total  area  of  the  dia- 
gram. In  the  turbine,  however,  it  is  different,  since  expansion 
can  be  caried  to  the  lower  back  pressure  line  within  the  turbine 
itself.  The  turbine  is  able  to  utilize  the  toe  of  the  diagram,  in- 
dicated by  the  shaded  portion  y,  in  addition  to  the  shaded  portion 
x,  while  the  engine  is  unable  to  turn  to  any  account  the  energy 
represented  by  the  toe  of  the  diagram. 

Theoretical  Gain  from  High  Vacuum. — An  idea  of  the  theo- 
retical gain  can  be  obtained  by  referring  to  a  few  calculations. 
Konrad  Anderson*  compares  power  values  for  steam  expanding 
from  60  and  200  pounds,  respectively,  and  finds  that  the  theoretical 
gain  in  running  condensing,  with  25  inches  vacuum,  over  run- 
ning non-condensing  is  nearly  100  per  cent  with  steam  at  60 
pounds  pressure,  and  50  per  cent  with  steam  at  200  pounds 
pressure.  If  the  vacuum  be  then  increased  to  28  inches,  the  gain 
with  the  60-pound  steam  will  be  about  22  per  cent  and  with  the 
200-pound  steam  about  18  per  cent.  This  shows  that  the  per- 

*Transactions  Institute  of  Engineers  and  Shipbuilders  of  Scotland,  1902. 


CONDENSING  APPARATUS  375 

centage  gain  from  running  condensing  is  more  with  low-pressure 
than  with  high-pressure  steam  and  that  the  gain  coming  from  the 
last  few  inches  of  vacuum  is  relatively  much  more  than  from  the 
first  few  inches. 

This  latter  fact  has  been  brought  out  in  a  striking  manner  by 
Ernest  N.  Janson.*  He  shows  that  with  the  initial  and  terminal 
pressures  in  the  same  ratio,  the  kinetic  energy  of  the  steam  in 
flowing  from  a  higher  to  a  lower  pressure  is  nearly  the  same, 
without  regard  to  what  the  initial  pressure  is.  For  example,  sup- 
posing the  initial  pressure  to  be  105  pounds  and  steam  to  expand 
to  one-third  this  pressure,  or  to  35  pounds,  he  finds  the  kinetic  en- 
ergy developed  by  the  steam  to  be  only  10  per  cent  more  than  when 
expanding  from  3  pounds  to  1  pound.  The  figures  are  as  fol- 
lows : — 

/>!         105 

—  — =  3 ;  velocity  =  2,050  ft.  per  sec. ;  H.  P.  per  Ib.  of 

ps         35 

steam  per  hour  =  0.033. 

P1        3 

—  =  —  =3;   vdocity=l,850   ft.   per  sec.;   H.    P.   per  Ib.   of 

/>,•        1 
steam  per  hour=0.027. 

Surface  Condenser  Plants. 

The  essential  features  of  a  condenser  plant  for  high  vacuums, 
as  built  by  the  Henry  R.  Worthington  Company,  are  shown  in  the 
diagram  Fig.  2.  Steam  enters  at  the  top  of  the  condenser  and  is 
distributed  over  the  tube  surface  by  baffle-plates,  while  the  con- 
densed steam  drops  down  into  the  hot  well  at  the  bottom,  where  it 
is  discharged  by  a  rotary  pump.  This  pump  requires  neither 
valves  nor  floats  and  is  not  subject  to  vapor  binding  as  are 
reciprocating  pumps.  The  capacity  of  the  pump  is  such  that  it 
runs  ahead  of  the  supply  so  that  the  suction  pipe  is  never  full ;  but 
the  discharge  pipe  is  always  full  and  the  water  presses  back 
against  the  pump.  As  long  as  the  latter  is  in  motion,  however, 

•Article  upon  steam  turbines  in  the  Journal  of  the  American  Society  of  Naval  En- 
gineers. 


r 


r 


i 

A 


CONDENSING  APPARATUS  377 

the  water  cannot  pass  back  into  the  condenser.  The  condenser  is 
fitted  with  an  air  cooler  which  is  frequently  applied  when  high 
vacuums  are  to  be  maintained.  This  is  simply  a  small  chamber 
containing  tubes  like  a  surface  condenser.  The  vapor  and  air 
from  the  condenser  pass  through  this  air  cooler,  where  they  are 
cooled  by  circulating  water  and  their  temperature  and  specific 
volume  thereby  reduced.  A  rotative  dry  vacuum  pump  exhausts 
the  air  and  vapor  from  the  air  cooler  and  maintains  a  high  vacuum. 
The  rotary  circulating  pump  driven  by  an  engine  is  used  for  the 
cooling  water.  It  is  usual  in  installations  of  this  kind  to  maintain 
practically  a  constant  supply  of  cooling  water,  sufficient  to  meet 
the  conditions  under  full  load. 

Wheeler  Condenser  and  Edwards  Air  Pump. — In  Fig.  3  is  an 
elevation  showing  one  of  the  500  Kw.  Curtis  turbines  at  the  New- 
port, R.  I.,  station  of  the  Massachusetts  Electric  Company.  This 
turbine  is  equipped  with  a  Wheeler  condenser  and  an  Edwards 
air  pump  made  by  the  Wheeler  Condenser  and  Engineering  Com- 
pany, New  York.  The  construction  of  this  pump  is  such  as  to 
make  one  of  the  simplest  possible  arangements  of  the  condensing 
apparatus.  Fig.  4  is  a  section  of  the  pump  cylinder.  It  has  no 
foot  valves,  which  require  a  pressure  in  the  condenser  somewhat 
above  that  in  the  pump  in  order  to  lift  them.  The  condensed 
steam  flows  continuously  by  gravity  from  the  condenser  into  the 
base  of  the  pump  and  is  there  dealt  with  mechanically  by  the 
conical  bucket  working  in  connection  with  a  base  of  similar  shape. 
Upon  the  descent  of  the  bucket  the  water  is  projected  at  a  high 
velocity  through  the  ports  into  the  working  barrel ;  the  plunger 
then  rises,  closing  the  ports,  and  sweeps  the  air  and  water  before 
it,  causing  them  to  escape  through  the  valve  at  the  top  of  the 
barrel.  The  elimination  of  the  foot  valves  in  this  pump  enables  a 
higher  vacuum  to  be  obtained  than  with  the  old  style  pump,  so 
that  27  or  28  inches  can  be  maintained  without  the  use  of  an  aux- 
iliary air  pump. 

As  indicated  in  Fig.  3  the  condenser  is  located  near  the  base  of 
the  turbine  and  in  front  of  it  are  the  Edwards  air  pump  and  a 
centrifugal  circulating  pump,  both  driven  by  electric  motor.  In 
this  plant,  as  in  others  arranged  according  to  modern  ideas,  the 
suction  sewer  and  discharge  sewer  for  the  circulating  water  are 


378 


STEAM  TURBINES 


nearly  on  the  same  level.  The  system  of  piping  leading  from  the 
suction  sewer,  through  the  circulating  pump  and  condenser,  and 
back  to  the  discharge  sewer,  thus  constitutes  a  closed  circuit,  one 
column  of  water  balancing  the  other.  The  sole  work  of  the  cir- 


Fig.  4.     Cross  Section  of  Edwards  Air  Pump. 

culating  pump,  therefore,  is  to  overcome  the  frictional  resistance 
of  the  water  flowing  through  the  piping  and  condenser  tubes. 

Curtis  Turbine  with  Condenser  in  Base. — In  Fig.  5  is  one 
of  the  5,000  Kw.  Curtis  turbine  units,  with  its  condenser  and 
other  auxiliaries,  installed  at  the  L-Street  station  of  the  Boston 
(Mass.)  Edison  Company.  In  this  case  the  condenser  is  built 
into  the  base  of  the  turbine  and  forms  a  part  of  the  unit, 
while  the  auxiliaries  are  on  the  same  floor  level  as  the  tur- 
bine itself,  where  they  are  more  accessible.  This  illustration 
gives  an  excellent  idea  of  the  quantity  of  apparatus  required 
to  keep  the  plant  in  operation,  since  the  feed  pumps,  heater, 
hot-well,  and  accumulator  for  supplying  the  hydraulic  pressure 


380 


STEAM  TURBINES 


that  must  be  maintained  under  the  step  bearing  of  the  turbine, 
etc.,  are  all  grouped  about  the  turbine,  in  addition  to  the  condenser 
auxiliaries.  The  lettered  parts  of  the  illustration  are  as  follows : 
A,  generator ;  B,  turbine ;  C,  condenser ;  D,  governor ;  E,  nozzles ; 
F ,  circulating  pump  ;  Gr  accumulator  for  step  bearing ;  H,  engine 
to  drive  circulating  pump ;  /_,  air  pump ;  K,  feed  pump ;  L,  heater ; 
M,  hot-well ;  N,  air-pump  engine. 


Fig.  6.     End  View  of  Turbine  and  Condenser  Shown  on  Opposite  Page. 

Westinghouse-P  arsons  Turbine  and  Alberger  Condenser. — 
Fig.  6  is  an  end  view  and  Fig.  7  a  plan  of  an  Alberger  surface 
condenser  and  apparatus  applied  to  a  5,000  Kw.  Westinghouse- 
Parsons  turbine.  The  Alberger  condenser  is  a  counter-current 
condenser  and  does  not  require  the  use  of  a  separate  air  cooler. 
The  exhaust  enters  at  the  bottom  and  passes  upward  over  the 
tubes.  The  cooling  water  enters  at  the  top  and  passes  downward, 


s 


CONDENSING  APPARATUS  383 

back  and  forth,  through  the  tubes.  The  air  and  vapors  rising  to 
the  top  of  the  condenser  are  therefore  cooled  by  the  in-coming 
cold  water  and  the  condensed  steam  which  trickles  down  into  the 
hot  well  is  heated  by  the  entering  steam  and  a  high  hot-well  tem- 
perature is  thus  maintained  without  difficulty.  A  two-stagfe 
vacuum  pump  is  employed,  one  cylinder  of  which  draws  the  air 
or  vapor  from  the  condenser  and  delivers  it  to  the  other  cylinder 
which,  in  turn,  forces  it  out  against  the  pressure  of  the  atmos- 
phere. Space  is  saved  and  the  apparatus  simplified  in  this  in- 
stallation by  using  the  same  engine  to  drive  both  the  circulating 
pump  and  the  vacuum  pump  and  it  will  be  noted  that  the  con- 
denser and  its  pumps  require  just  about  the  same  area  as  the  tur- 
bine itself,  while  by  setting  the  condensing  apparatus  in  a  pit  it 
rises  only  to  the  top  of  the  turbine. 

Condensing  Apparatus  Underneath  the  Turbine. — It  is  advo- 
cated by  many  that  the  proper  place  for  the  condenser  and  aux- 
iliaries is  on  the  same  floor  as  the  turbine,  where  they  will  be 
likely  to  receive  better  attention  than  if  placed  in  a  basement  be- 
tween the  foundation  piers  or  underneath  the  turbine.  In  view  of 
the  fact  that  a  less  massive  foundation  is  required  for  a  turbine 
than  for  an  engine,  however,  it  is  possible  to  house  the  condensing 
apparatus  perfectly  under  the  turbine,  as  already  explained  in 
Chapter  XVII.  Such  a  plan,  which  in  this  instance  is  an  ex- 
ceedingly neat  arrangement,  has  been  followed  in  the  installation 
at  a  power  house  of  the  Glasgow  (Scotland)  Corporation,  Fig.  8. 
This  is  a  Willans  &  Robinson-Parsons  turbine.  The  condenser, 
circulating  pump  and  Edwards  air  pump  are  clearly  shown,  the 
latter  two  being  driven  by  electric  motor.  The  lettered  parts  are : 
A,  condenser ;  B,  air  pump ;  C,  motor  for  pump ;  D,  circulating 
pump ;  E,  motor. 

Parsons  Vacuum  Augmenter. — A  novel  arrangement  for  main- 
taining a  high  vacuum  is  the  vacuum  augmenter  employed  by 
Parsons  in  England  with  considerable  success.  He  uses  air 
pumps  as  shown  in  Fig.  9  placed  below  the  level  of  the  condenser 
and  in  addition  to  the  usual  pipe  connection  between  the  air 
pumps  and  the  condenser  there  is  a  small  pipe  leading  to  the 
auxiliary  condenser,  generally  having  about  one  twentieth  the 
cooling  surface  of  the  main  condenser.  In  a  contracted  portion  of 


384 


STEAM  TURBINES 


this  auxiliary  pipe  is  a  steam  nozzle  which  discharges  a  jet  of 
steam  that  acts  similar  to  the  jet  of  an  injector;  this  jet  draws 
nearly  all  the  residual  air  and  vapor  from  the  condenser  and  de- 
livers it  to  the  air  pumps.  The  main  pipe  leading  to  the  air  pump 
is  so  curved  as  to  form  an  air  seal  which  prevents  the  air  and 
vapor  from  returning  to  the  condenser.  With  this  arrangement 
there  need  be  a  vacuum  in  the  air  pumps  of  only  about  26  inches, 
while  the  vacuum  augmenter  will  increase  the  vacuum  in  the  con- 
denser to  27  or  28  inches.  Mr.  Parsons  states  that  the  quantity 
of  steam  required  for  the  steam  jet  is  about  \l/2  per  cent  of  that 


Fig.    9.     Parsons   Vacuum   Augmenter. 

Jet  and  Injector  Condensers. 


used  by  the  turbine  at  full  load,  and  this,  together  with  the  air 
extracted,  is  cooled  by  the  auxiliary  condenser. 

Surface  vs.  Jet  Condensers. — The  surface  condenser  has  come 
into  extensive  use  with  the  steam  turbine  because  the  steam  dis- 
charging from  a  turbine  is  entirely  free  from  oil  and  if  collected 
and  condensed  can  be  used  over  and  over  in  the  boilers.  The  feed 
water  leaves  the  hot  well  of  a  surface  condenser  operating  at  high 
vacuum  at  nearly  100  degrees  F.,  and  passes  through  a  heater 
where  the  temperature  is  raised  still  further  by  steam  from  the 
auxiliaries.  In  jet  and  injector  condensers  the  condensed  steam 
passes  off  with  the  injection  water,  which  is  at  a  temperature  of 


CONDENSING  APPARA  TVS  385 

80  or  90  degrees,  and  when  part  of  this  is  used  for  boiler  feed 
there  is  a  loss  of  some  10  or  20  heat  units  per  pound,  as  compared 
with  the  surface  condenser.  This  is  so  slight,  however,  that  it 
does  not  pay  to  install  a  surface  condenser  and  attending  apparatus 
on  the  score  of  heat  saved.  The  jet  type  is  cheaper,  simpler  and 
works  as  well  or  better. 

In  localities  where  the  available  water  supply  contains  sulphate 
of  lime,  acid,  grease,  or  other  harmful  impurities,  or  where  the 
cost  of  pure  water  is  high,  the  surface  condenser  should  probably 
be  given  the  preference,  though  if  it  is  merely  the  cost  of  the 
water  that  is  at  stake,  the  problem  should  be  gone  into  very  care- 
fully before  deciding.* 

Injector  Condenser. — A  Bulkley  injector  condenser  was  in- 
stalled by  Geo.  I.  Rockwood  in  connection  with  a  Westinghouse- 
Parsons  turbine  at  Providence,  R.  I.  The  injection  water  is  ele- 
vated into  a  vertical  tank  30  inches  square  by  15  feet  deep,  in 
which  the  water  level  is  maintained  6  inches  below  the  water  inlet 
nozzle  of  the  condenser.  The  injection  pipe  takes  the  water  from 
near  the  bottom  of  the  tank.  The  air  entrained  with  the  water 
rises  to  the  top  of  the  tank  and  is  largely  eliminated  from  the 
injection  water  entering  the  condenser.  The  flow  of  the  injection 
water  through  the  throat  of  the  condenser  is  what  constitutes  the 
air  pump,  and  it  is  found  to  be  the  only  air  pump  needed,  since  a 
vacuum  of  28*/2  inches  has  been  maintained,  regardless  of  whether 
steam  is  passing  through  the  turbine  or  not. 

Jet  Condenser. — A  jet  condenser,  which  is  a  modification  of  the 


*In  a  paper  before  the  A.  S.  M.  E.,  December,  1904,  Geo.  I.  Rockwood  contends  that 
the  injector  condenser  would  seem  to  bar  out  all  other  condenser  systems  in  situations 
where  the  water  is  pure.  He  gives  figures  to  show  that  it  does  not  pay  to  install  a 
surface  condenser  simply  to  save  paying  city  rates  for  boiler  feed  water.  His  esti- 
mate for  the  cost  of  a  high-vacuum  surface  condenser  outfit  is  from  $7  to  $10  per 
kilowatt  and  of  a  jet  or  barometric  condenser  system  from  $5  to  $6  per  kilowatt. 

In  a  paper  before  the  American  Railway  Mechanical  and  Electrical  Association,  1905, 
Fred  N.  Bushnell  writes:  "In  cases  where  the  cost  of  feed  water  is  a  material  factor 
in  the  cost  of  power,  or  where  it  contains  a  large  percentage  of  calcium  or  magnesium 
carbonate,  or  other  scale-forming  materials,  there  will  be  great  advantage  in  using 
a  surface  condenser  on  account  of  the  pure  distilled  water  returned  to  the  boilers, 
but  where  these  conditions  do  not  exist  it  will  frequently  be  found  practicable  to  use 
some  simpler  form  of  condensing  apparatus  such,  for  example,  as  the  injector  or 
barometric  type  of  jet  condensers.  These  types  of  condensers  offer  very  great  ad- 
vantages over  the  surface  condenser  in  the  matter  of  lower  first  cost,  space  occupied, 
greater  simplicity,  and  less  cost  of  maintenance.  Up  to  this  time  they  have  not  been 
very  generally  used,  but  there  seems  to  be  no  good  reason  why  they  should  not  work 
as  satisfactorily  in  connection  with  steam  turbines  as  with  reciprocating  engines." 


386 


STEAM  TURBINES 


injector  condenser,  is  made  by  the  Worthington  company.  As  in 
the  injector  type,  the  condenser  proper  is  placed  about  30  feet 
above  the  hot  well  and  the  water  falling,  through  the  action  of 
gravity,  creates  the  vacuum.  There  is  no  contracted  throat  to 
this  condenser,  however,  and  the  water  is  sprayed  into  the  head, 
where  it  becomes  intimately  mingled  with  the  steam  before  dis- 
charging through  the  vertical  pipe.  Fig.  10  shows  a  section  of 
the  condenser  head.  An  air  cooler  and  a  dry  vacuum  pump  are 


OPEHIHG  TO  TAIL  PIPE 


Fig.    10.     Section    of   Worthington 
Jet     Condenser. 

employed,  such  as  used  with  surface  condensers,  and  any  air  that 
accumulates  in  the  condenser  head,  where  the  steam  is  condensed 
by  the  spray  of  the  water,  is  removed  by  the  pump. 

An  interesting  application  of  a  jet  condenser  to  a  Parsons  tur- 
bine is  shown  in  Fig.  11.  The  plan  is  here  adopted  of  sub- 
stituting a  centrifugal  pump  for  the  usual  barometric  column, 
enabling  the  condenser  to  be  placed  under  the  turbine.  The  ex- 
haust steam  is  led  through  a  pipe,  A,  and  a  gate  valve,  B,  into  the 
condensing  chamber,  C;  and  there,  it  is  condensed  by  a  jet  and 
flows  into  the  opening  of  a  centrifugal  pump,  which  is  driven  by 
a  belt  from  the  pulley  on  the  extended  shaft  of  the  turbine. 


CONDENSING  APPARATUS 


387 


PULLEY  CASE 


Fig.   11.     Novel  Arrangement  of  Jet  Condenser  with   Centrifugal   Pump. 

There  is  a  check,  D,  in  the  discharge  from  the  pump,  and  this 
discharge  is  also  sealed  by  the  outgoing  water.  At  the  top  of  the 
condensing  chamber  is  attached  the  usual  dry  air  pump  connec- 
tion. 

Data  in  Regard  to  the  Performance  of  Condensers  Operating  Under 

High  Vacuum. 

Quantity  of  Cooling  Water  and  Area  of  Condensing  Surface. — 
An  ordinary  surface  condenser  giving  26  inches  vacuum,  with 
cooling  water  at  70  degrees,  requires  about  30  pounds  of  cooling 
water  per  pound  of  steam  condensed  and  will  condense  about  10 
pounds  of  steam  per  hour  per  square  foot  of  surface.  A  surface 
condenser  to  operate  with  28  inches  vacuum  will  require  more 
than  twice  this  amount  of  cooling  water,  or  say,  70  pounds  per 
pound  of  steam  (one  manufacturer  estimates  as  high  as  85 


388  STEAM  TURBINES 

pounds)  and  will  condense  about  5  pounds  of  steam  per  square 
foot  of  cooling  surface.  A  common  allowance  in  turbine  work 
is  4  square  feet  of  cooling  surface  per  kilowatt.  If  the  tempera- 
ture of  the  cooling  water  is  above  70  degrees  the  weight  of 
water  will  have  to  be  increased,  sometimes  very  largely.  A  less 
quantity  of  injection  water  is  required  for  jet  or  barometric  con- 
densers than  for  surface  condensers.  One  manufacturer  of  the 
barometric  type  has  furnished  the  author  with  the  following  fig- 
ures for  a  28-inch  vacuum: — 

Injection  at  40  deg.,  26  Ib.  per  Ib.  steam 
«         «   CQ     t(     29   "     "     "       " 

"  60     "     35    "     

"  70     "     50   

A  common  allowance  is  60  pounds  per  pound  of  steam  for 
water  at  70  degrees. 

Tests  on  Condensers. — There  are  no  data  yet  available,  at 
least  to  the  public,  in  regard  to  the  performance  of  high-vacuum 
condensers,  by  which  the  relations  between  the  several  elements 
entering  into  the  calculation  of  the  quantity  of  cooling  water,  area 
of  tube  surface,  etc.,  can  be  established,  and  results  for  the  present 
must  be  more  or  less  empirical. 

Before  calculations  of  condenser  performance  can  be  made,  the 
initial  and  final  temperatures  of  the  steam  and  cooling  water  must 
be  known.  The  following  figures  were  given  to  the  author  at 
the  works  of  the  B.  F.  Goodrich  Company,  Akron,  O.,  where 
Westinghouse-Parsons  turbines  are  installed. 

Test  1.  Barometer  30  inches;  vacuum  28.26;  temperature 
steam  in  condenser  99.5 ;  temperature  hot  well  99.05 ;  initial  tem- 
perature injection  water  54  degrees;  final  temperature  injection 
water  73  degrees. 

Test  2.  Barometer  30  inches;  vacuum  28.61;  temperature 
steam  in  condenser  100.5;  temperature  hot  well  100.5;  initial  tem- 
perature injection  water  57.9;  final  temperature  injection  water 
72.5. 

The  data  in  regard  to  the  quantity  of  water  used  and  the  power 
developed  by  the  turbine  during  the  tests  were  not  sufficiently 
accurate  to  denote  exact  results,  but  they  indicated  about  60 
pounds  cooling  water  per  pound  of  steam  condensed. 


CONDENSING  APPARA  TUS  389 

The  following  results  are  from  tests  upon  a  surface  condenser 
with  Edwards  air  pump,  in  connection  with  a  Curtis  turbine  which 
was  running  at  a  very  light  load : — 

Test  1.  Test  2.  Test  3. 


Vacuum,  inches,  27.89  27.64  27.92 

Area  cooling  surface,  sq.  ft.,  2,700.  2,700.  2,700. 
Temperatures,  degrees  F. : 

Steam  in  condenser,  103.  107.  102. 

Hot  well,  93.5  98.  93.5 

Cooling  water,  initial,  75.  75.  75. 

Cooling  water,  final,  96.  98.  93. 

Weight  steam  per  hour,  lb.,  270.66  309.09  317.19 

Weight  cooling  water  per  hour,  13,560.  14,094.9  19,634.2 

Ratio,  water  to  steam,  50.1  45.6  61.9 

These  three  tests  show  a  condition  that  is  seldom  met  with  in 
practice ;  viz.,  a  final  temperature  of  the  cooling  water  equal  to  or 
higher  than  the  hot-well  temperature.  This  was  attained  because 
of  the  small  quantity  of  steam  condensed  per  square  foot  of  cool- 
ing surface.  Under  ordinary  conditions  the  final  temperature  of 
the  cooling  water  will  be  from  10  to  25  degrees  below  the  hot- 
well  temperature.  In  counter-current  condensers  the  tempera- 
ture will  be  higher  than  in  the  parallel  flow  type.  The  final  tem- 
perature is  also  dependent  upon  the  quantity  of  cooling  water 
forced  through  the  condenser  tubes  and  upon  the  area  of  the  tube 
surface. 

Condenser  Calculations. — The  following  simple  example  shows 
the  method  of  calculating  the  weight  of  cooling  or  injection  water 
when  the  temperatures  are  known.  No  allowance  is  here  made 
for  the  efficiency  of  the  condenser,  which  must  be  determined  by 
experiment,  but  the  example  will  explain  why  it  is  so  difficult  to 
maintain  a  high  vacuum. 

Example: — The  temperature  of  the  steam  in  a  condenser  at 
28  inches  vacuum  is  about  100  degrees,  and  its  total  heat  per 
pound  is  found  from  the  steam  tables  to  be  1,112  units.  By  the 
use  of  a  dry  vacuum  pump  it  is  possible  to  secure  a  hot  well  tem- 
perature of  98  degrees,  and  the  heat  in  each  pound  of  condensed 
steam  is  therefore  98  —  32  =  66  units.  Hence  there  are  1,112  — 


390  STEAM  TURBINES 

66  =  1,046  heat  units  given  up  to  the  cooling  water  per  pound  of 
steam  condensed. 

With  cooling  water  at  70  degrees  initial  and  100  degrees  final 
temperature  (the  latter  equal  to  the  temperature  of  the  condensed 
steam),  the  water  will  have  risen  30  degrees  and  gained  30  heat 
units  per  pound. 

But  the  heat  lost  by  the  steam  was  1,046  heat  units  per  pound. 
Hence,  l,046-f-30=34.9  for  the  ratio  of  cooling  water  to  con- 
densed steam. 

In  practice  no  condenser  can  work  with  so  high  an  efficiency  as 
in  this  case,  where  the  cooling  water  takes  up  all  the  heat  of  con- 
densation of  the  steam  and  leaves  the  condenser  with  the  tempera- 
ture of  the  condensed  steam.  Under  normal  conditions  the  final 
temperature  of  the  cooling  water  will  range  from  10  to  25  degrees 
below  the  temperature  of  the  condensed  steam,  as  previously 
stated.  Let  us  assume  it  to  be  15  degrees  below.  Then,  with  cool- 
ing water  at  70  degrees  initial  and  85  degrees  final  temperature, 
we  have : — 

Heat  units  absorbed  r~r  pound  of  water=85 — 70=15;  and 
1,046-^15=69.7,  for  the  ratio  of  cooling  water  to  condensed  steam, 
or  double  what  it  was  before. 

In  winter  time,  when  the  initial  temperature  of  the  cooling 
water  would  be  about  40  degrees,  we  should  have,  assuming  a 
final  temperature  of  85  degrees,  85 — 40=45  and  1,046-^45=23.2 
for  the  ratio.  As  a  matter  of  fact  it  would  be  more  likely  that  the 
plant  would  be  operated  with  as  much  cooling  water  in  winter  as 
in  summer  to  derive  the  benefit  of  the  higher  vacuum  that  would 
be  secured;  and  in  this  case  the  final  temperature  would  drop  to, 
say  25  degrees  below  the  temperature  of  the  condensed  steam,  as 
was  the  case  in  the  tests  upon  the  Goodrich  plant  previously 
quoted. 

On  the  other  hand,  if  cooling  towers  were  employed  and  the 
temperature  of  the  cooling  water  rose  to  80  degrees  or  more  in 
the  summer  time,  it  is  evident  that  the  square  feet  of  cooling  sur- 
face of  the  condenser  should  be  on  a  liberal  basis,  in  order  to 
secure  as  high  a  final  temperature  of  the  cooling  water  as  possible ; 
otherwise  the  quantity  of  water  to  be  circulated  might  be  almost 
prohibitive  in  amount.  Even  under  favorable  conditions  the  prob- 


CONDENSING  APPARATUS  391 

lem  of  maintaining  a  high  vacuum  in  connection  with  a  cooling 
tower  is  doubtful  of  solution. 

Power  to  Operate  Auxiliaries.* — The  plant  of  the  Citizens' 
Light,  Heat  and  Power  Company,  Johnstown,  Pa.,  is  partially 
equipped  with  turbines  and  Weiss  jet  condensers  of  the  barometric 
type.  The  condenser  auxiliaries  are  driven  from  a  single  steam 
cylinder — that  of  the  rotative  air  pump — and  by  indicating  this 
cylinder  at  normal  speeds  the  total  power  in-put  was  obtained. 
The  results  were  reduced  to  percentages  of  the  turbine  output  and 
showed  that  at  less  than  one  quarter  load  the  total  power  con- 
sumption was  less  than  five  per  cent  of  the  sustained  output,  and 
that  it  progressively  decreased  to  2^  per  cent  at  full  load.f 

J.  R.  Bibbins  states  in  a  paper,  "Steam  Turbine  Power  Plants," 
that  observations  on  the  2,000  Kw.  turbine  at  the  St.  Louis  Expo- 
sition indicated  the  following  results :  At  full  load  the  total  power 
in-put  to  the  auxiliaries  was  7  per  cent;  approximately  three 
fourths  required  for  the  circulating  pump,  two  ninths  for  the  air 
pump,  and  one  fourteenth  for  the  hot  well  pump.  This  7  per  cent 
represents  gross  power  and  includes  transformer  and  motor  losses. 
The  cooling  water,  however,  was  exceptionally  warm,  being  fre- 


*In  the  Journal  of  Electricity,  Power  and  Gas  for  March,  1905,  was  the  first  of 
several  extended  articles  by  Charles  C.  Moore  &  Co.,  incorporated  engineers  at  San 
Francisco,  Cal.,  upon  the  estimated  power  for  operating  condenser  auxiliaries  under 
above  conditions.  The  following  quotation  is  taken  from  the  introduction:  "The 
quantity  of  circulating  water  required  for  high-vacuum  condensing  plants  must  be  in- 
creased from  the  old  standard  of  from  25  to  30  pounds  to  from  40  to  60  pounds  of 
water  per  pound  of  steam  condensed  for  moderate  temperatures,  and  from  60  to  100 
pounds  of  water  per  pound  of  steam  when  used  at  the  higher  temperatures  common 
to  cooling  tower  practice.  In  cases  of  excessive  head  or  quantity  of  circulating  water 
the  bulk  of  the  power  required  by  auxiliaries  is  due  to  the  circulating  pump.  The 
range  of  power  for  this  purpose  varies  so  widely  that  the  older  method  of  assuming 
a  given  type  of  plant  requiring  5,  10,  or  15  per  cent  of  the  total  steam  consumption 
to  drive  auxiliaries  is  entirely  in  error  without  an  accompanying  statement  defining 
conditions  under  which  circulating  water  is  pumped.  The  power  to  drive  the  air 
pump  is  dependent  somewhat  upon  the  vacuum,  but  particularly  upon  the  air  leakage 
into  the  condensing  system.  It  was  for  some  time  assumed  that  the  work  of  the  air 
pump  corresponded  to  removing  the  air  which  entered  the  boiler  in  solution  in  feed 
water.  As  a  matter  of  fact,  handling  the  air  in  solution  is  the  smallest  portion  of 
work  done  by  air  pump,  the  leakage  through  piping,  pipe  joints,  pores  of  castings, 
stuffing  boxes,  etc.,  imposing  the  greatest  duty,  the  total  quantity  of  air  to  be  handled 
ranging  from  ten  or  fifteen  to  thirty  or  forty  times  the  air  dissolved  in  ordinary 
water.  The  actual  power  to  drive  the  air  pump  should  in  good  practice  be  less  than 
.00018  indicated  horse-power  in  the  air  pump  cylinder  per  pound  of  exhaust  steam 
per  hour.  As  the  amount  of  power  necessary  to  drive  the  air  pump  is  a  comparatively 
small  portion  of  the  total  power  for  auxiliaries  a  slight  error  in  this  quantity  will  not 
largely  affect  the  final  result." 

^Power,  February,  1905. 


392  STEAM  TURBINES 

quently  85  degrees,  which  made  it  difficult  to  obtain  the  high 
vacuum  desired.* 

The  following  tests  upon  the  auxiliaries  of  the  5,000  Kw.  unit 
of  the  Boston  Edison  Company,  shown  in  Fig.  5,  were  reported  in 
the  report  of  the  turbine  committee  of  the  National  Electric  Light 
Association  for  1905  : — 

Test  1.  Test  2.  Test  3. 


Kilowatts  on  turbine,  2,713.  3,410.  4,758. 

Vacuum,  28.4  28.7  28.6 

Barometer,  29.53  29.95  29.96 


Horse-power  Developed. 

Boiler  feed  pump,  13.9  23.7  27.4 

Circulating  pump,  69.1  69.1  69.1 

Dry  vacuum  pump,  24.3  23.2  23.8 

Step  bearing  pump,  6.4  5.8  5.6 

Wet  vacuum  pump,  8.6  9.2  9.8 


Totals,  122.3  131.  135.7 

Per  cent  power  of  auxiliaries  to  power 

of  turbine,  3.4  2.9  2.1 

Per  cent  water  used  by  auxiliaries  to 

that  used  by  turbine,  8.4  7.4  5.7 


* American   Street  Railway  Association,   1904. 


CHAPTER  XX 
THE  STATUS  OF  THE  MARINE  TURBINE. 

Early  History. — Most  of  the  turbines  applied  to  the  propul- 
sion of  vessels  have  been  of  the  Parsons  type,  although  some  work 
of  this  character  has  been  done  both  by  Rateau  and  Curtis.  In 
1894  the  Parsons  Marine  Steam  Turbine  Company,  Ltd., 
Wallsend-on-Tyne,  England,  was  formed  and  the  experimental 
boat  Turbinia  constructed.  Her  dimensions  were  100  feet  beam, 
3  feet  draft  and  44  tons  displacement.  There  were  three 
separate  turbines — a  high-,  an  intermediate-,  and  a  low-pressure, 
each  driving  a  screw  shaft  and  on  each  shaft  were  keyed  three 
propellers  of  small  diameter.  The  turbines  were  rated  at  2,000 
horse-power  and  the  boat  attained  a  speed  of  over  34  knots. 

Various  other  high  speed  boats  were  built  during  the  next  five 
or  six  years.  Two  of  these,  the  Viper  and  Cobra,  high  speed 
torpedo  boat  destroyers,  were  lost  at  sea  and  turbine  propulsion 
received  a  serious  setback.  An  organization  was  finally  effected, 
however,  which  included  the  shipbuilding  firm  of  Messrs.  Denny, 
the  Hon.  Charles  A.  Parsons  and  Capt.  John  Williamson,  which 
resulted  in  the  first  turbine  steamer,  King  Edward,  in  1901,  for 
service  on  the  Clyde. 

The  First  Turbine  Steamer. — The  King  Edward  is  a  boat  250 
feet  long,  30  feet  beam  with  6  feet  draft.  The  arrangement  of  the 
machinery  is  practically  the  same  as  has  been  used  in  all  the  more 
recent  vessels,  including  the  ocean  liners,  fitted  with  Parsons 
turbines.  There  are  three  separate  turbines  driving  three  screw 
shafts.  The  high-pressure  turbine  is  placed  on  the  center  shaft 
and  the  two  low-pressure  turbines  each  drive  one  of  the  outer 
shafts.  Inside  the  exhaust  ends  of  each  of  the  latter  are  placed 
the  two  astern  turbines  which  rotate  as  one  piece  with  the  low- 
pressure  motors  and  when  in  operation  reverse  the  direction  of 
rotation  of  the  low-pressure  motors  and  outside  shafts. 

In  ordinary  going  ahead  steam  from  the  boilers  is  admitted  to 
the  high-pressure  turbine  and  after  expanding  about  5  times 
passes  to  the  low-pressure  turbines  and  is  again  expanded  in 


394  STEAM  TURBINES 

them  about  25  times  and  then  passes  to  the  condensers,  the  total 
expansion  ratio  being  about  125  as  compared  with  from  8  to  1C 
usual  in  triple  expansion  reciprocating  engines  of  the  marine  type. 
At  20  knots  the  speed  of  the  center  shaft  is  700  and  of  the  two 
outer  shafts  1,000  per  minute. 

When  maneuvering  in  or  out  of  harbor  the  outer  shafts  only 
are  used  and  the  steam  is  admitted  by  suitable  valves  directly  into 
the  low-pressure  motors  or  into  the  reversing  motors,  for  going 
ahead  or  astern.  The  high-pressure  turbine  under  these  circum- 
stances revolves  idly,  its  steam  admission  valve  being  closed  and 
its  connection  with  the  low-pressure  turbines  being  also  closed  by 
non-return  valves. 

Later  Turbine  Boats. — Following  the  King  Edward,  and  a  later 
boat  for  the  same  line,  the  Queen  Alexandria,  has  come  a  long 
list  of  other  turbine  vessels,  notably  a  fleet  of  18  cross-Channel 
boats  built  or  building,  to  ply  between  Dover  and  Calais.  Again, 
on  the  Heysham  line  running  between  Great  Britain  and  Ireland, 
turbine  vessels  have  been  in  successful  operation.  In  1904  the 
third-class  turbine  cruiser  Amethyst  was  built  for  the  British 
Admiralty.  She  is  360  feet  in  length  and  of  3,000  tons  displace- 
ment. Three  other  engine-driven  cruisers  of  the  same  size  were 
built  simultaneously,  one  of  which,  the  Topaz,  was  selected  for  a 
series  of  competitive  trials  with  the  Amethyst. 

The  contract  speed  of  the  vessels  was  21)4  knots,  and  the  results 
showed  that  at  all  speeds  above  14*^  knots  the  turbine  vessel  was 
the  more  economical,  at  18  knots  the  turbine  was  15  per  cent  more 
economical,  at  20^  knots  31  per  cent,  at  22.1  knots  36  per  cent, 
and  at  full  power  in  each  vessel  the  Amethyst  showed  42  per  cent 
more  power  than  required  by  contract  on  the  coal  allowed ;  while 
the  Amethyst  reached  23.6  knots  on  the  specified  coal  and  the 
Topaz  only  22.1  knots.  In  other  words,  the  Amethyst  has  a  radius 
of  action  at  20  knots  speed  of  3,600  nautical  miles,  while  her  sister 
vessels  with  ordinary  engines  can  only  steam  2,000  miles  at  the 
same  speed. 

The  success  of  the  Amethyst  led  British  naval  constructors  to 
advocate  the  turbine  for  larger  vessels  and  the  activity  of  the 
admiralty  following  the  Russo-Japanese  war  culminated  in  the 
construction  of  the  powerful  battleship  Dreadnought.  This  ship, 


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396  STEAM  TURBINES 

which  is  not  only  larger  and  carries  a  heavier  armament  than  any 
battleship  afloat,  is  remarkable  because  it  is  the  first  battleship  to 
be  driven  by  turbines.  Under  trial  the  turbines  developed  28,000 
horse-power  and  propelled  the  vessel  at  an  average  speed  of  21^ 
knots  during  a  trial  of  eight  hours,  acquiring  a  maximum  speed 
of  22^4  knots.  The  turbines  are  so  free  from  vibration  that  the 
ship  makes  the  steadiest  possible  gun  platform  for  a  floating 
battery. 

Atlantic  Liners  Fitted  with  Turbines. 

Description  of  the  Steamer  Victorian. — The  Allan  liners  Vir- 
ginian and  Victorian  started  to  ply  between  Liverpool  and 
Canada  in  the  summer  of  1905.  In  April,  1905,  Commander  A.  D. 
Canaga,  United  States  Navy,  was  detailed  to  make  the  trip  to 
Europe  and  return  on  the  turbine  steamer  Victorian  and  report 
the  results  of  his  observation  to  the  department*  There  unfor- 
tunately is  no  similar  vessel  of  the  same  line  propelled  by  recipro- 
cating engines  with  which  a  direct  comparison  can  be  made,  but 
certain  points  brought  out  by  Commander  Canaga  will  be  of  in- 
terest. Figs.  1,  2,  and  3  are  reproduced  from  his  report  show- 
ing the  arrangement  of  turbines  in  this  vessel  which  is  like  that 
usually  adopted  for  the  Parsons'  apparatus,  and  is  practically  the 
same  as  already  described  in  connection  with  the  King  Edward. 
The  steam  from  the  boilers  is  led  into  the  engine  room  through 
two  12-inch  pipes,  uniting  in  the  throttle  valve  at  the  working 
platform.  From  the  throttle  valve  steam  is  led  through  two 
12-inch  pipes  to  the  high-pressure  turbine.  When  in  free  route 
the  steam  is  passed  through  the  high-pressure  turbine  where  it 
spreads,  half  going  to  the  starboard  and  half  to  the  port  turbine 
through  the  receiver  pipes,  and  thence  through  exhaust  pipes  to 
the  main  condensers.  In  maneuvering,  the  main  throttle  is  closed 
and  steam  admitted  to  the  maneuvering  valves,  Fig.  3,  one  for 
each  low-pressure  turbine.  These  are  simple  slide  valves  which 
when  placed  at  the  upper  end  of  their  stroke  admit  live  steam  to 
the  forward  end  of  the  low-pressure  turbine,  when  at  the  bottom 
of  their  stroke  admit  live  steam  to  the  backing  turbine,  and  when 


* 'Journal  of  the  American  Society  of  Naval  Engineers,  August,  1905. 


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398  STEAM  TURBINES 

in  mid-position  shut  the  steam  from  both  the  ahead  and  backing 
turbines. 

To  prevent  the  steam  blowing  off  into  the  high-pressure  tur- 
bine when  maneuvering,  non-return  valves  are  fitted  in  the  receiver 
pipes  between  the  high-pressure  and  low-pressure  turbines,  as 
shown  in  Fig.  2.  These  valves  are  automatic,  opening  or  closing 
as  the  live  steam  is  admitted  to  the  high-pressure  or  low-pressure 
turbines. 

Comments  on  the  Operation  of  the  Victorian. — At  the  forward 
end  of  each  turbine  shaft  is  fitted  a  safety  governor,  which  in  case 
of  accident  closes  the  main  throttle  valve.  These  governors  serve 
another  purpose,  also,  by  indicating  whether  the  turbines  are  at 
rest  or  in  motion,  since  from  the  working  platform  they  are  the 
only  visible  moving  parts.  The  commander  reports  that  the  tur- 
bines are  easily  and  quickly  handled  and  that  the  minor  mishaps 
and  annoyances  met  with  in  reciprocating  engines  are  absent.  He 
notes  a  pleasing  absence  of  vibration  and  of  racing  in  high  seas. 
Against  this  immunity  from  racing,  however,  must  be  set  the  lack 
of  holding  power  of  the  small  screws  with  which  turbine  vessels 
must  be  equipped.  It  was  observed  that  the  influence  of  head 
winds  and  heavy  seas  reduced  the  vessel's  speed  considerably  more 
than  would  have  been  the  case  with  the  large  propellers  used  with 
reciprocating  engines. 

Attention  is  called  to  the  fact  that  heating  of  bearings  is  a  more 
serious  matter  with  turbine  machinery  than  with  reciprocating 
engines  as  any  unusual  wear  would  cause  interference  between 
the  rotating  and  stationary  blades,  and  on  one  of  the  trips  the 
Victorian  was  delayed  about  29  hours  owing  to  some  grit  that  got 
into  one  of  the  bearings  and  necessitated  overhauling,  after  which, 
however,  there  was  no  trouble.  There  was  also  considerable  diffi- 
culty from  priming  of  the  boilers,  but  with  consequences  less 
serious  than  in  the  case  of  reciprocating  engines. 

Turbine  Boats  of  the  Cunard  Line. 

The  Carmania*  was  built  and  equipped  by  Messrs.  John  Brown 
&  Co.,  Ltd.,  Gydebank.  This  vessel  is  one  of  the  seven  or  eight 


*Taken  in  part  from  London  Engineering,  December  1,  1905. 


STATUS  OF  THE  MARINE  TURBINE  399 

largest  of  the  world's  ships,  having  the  following  dimensions: 
length  between  perpendiculars,  650  feet;  length  over  all,  672  feet, 
6  inches;  breadth,  moulded,  72  feet;  depth,  moulded,  52  feet; 
gross  register  tonnage,  19,524  tons;  draft,  in  working  condition, 
33  feet,  3%  inches ;  displacement  at  this  draft,  30,918  tons.  She 
was  designed  for  a  speed  of  19  knots. 

The  Turbines  are  of  the  Parsons  type  arranged  in  the  usual 
manner,  with  the  high-pressure  turbine  on  the  center  screw  shaft 
and  with  one  low-pressure  turbine  on  each  of  the  outside  shafts. 
The  arrangement  of  the  turbines  and  auxiliaries  is  shown  in 
Figs.  4  and  5.  In  normal  working  conditions  the  regulating  valves 
admit  steam  to  the  high-pressure  turbine  and  the  steam  passes 
through  it  to  the  low-pressure  turbines  and  thence  to  the  con- 
denser. For  maneuvering  purposes  a  large,  non-return  valve, 
worked  by  a  steam  and  hydraulic  engine  controlled  from  the  start- 
ing platform,  closes  the  connection  between  the  high-pressure  and 
each  low-pressure  turbine.  This  valve  will  close  automatically  as 
soon  as  a  prescribed  pressure  is  obtained  in  the  low-pressure 
casing.  Each  of  the  low-pressure  turbines  is  then  manipulated  as 
an  independent  unit.  Special  maneuvering  valves  are  fitted  to  each 
of  the  low-pressure  and  astern  turbines,  which  allow  of  steam  being 
admitted  to  either  the  ahead  or  astern  turbines  by  a  single  move- 
ment of  a  hand  lever.  The  turbine  is  fitted  with  a  governor  which 
operates  when  there  is  any  marked  increase  in  the  number  of  revo- 
lutions. Oil  is  supplied  to  the  bearings  under  considerable  pres- 
sure. It  then  flows  to  a  cooling  tank  fitted  with  copper  coils 
through  which  water  is  circulated  and  after  passing  through  a 
system  of  oil  filters  is  again  delivered  to  the  bearings. 

The  Turbine  Glands. — The  gland  for  the  shaft  passing  through 
the  end  of  the  turbine  is  rendered  steam-tight  by  an  improved 
design,  Fig.  6.  In  the  Parsons  marine  turbine  of  smaller  sizes  the 
leakage  of  steam  from  the  high-pressure  turbine  and  the  ingress  of 
air  into  the  low-pressure  turbine  has  been  prevented  by  the  Parsons 
ring-and-groove  type  of  gland.  This  consists  of  a  series  of 
grooves  turned  in  the  spindles  into  which  bronze  rings  are  fitted, 
the  whole  rotating  in  a  truly  bored  cylindrical  gland.  When  the 
difference  in  pressure  between  the  inside  of  the  turbine  and  the 
atmosphere  is  great  the  side  pressure  of  these  rings  is  very  con 


STATUS  OF  THE  MARINE  TURBINE  401 

siderable  and  in  order  to  distribute  this  pressure  as  equally  as 
possible  the  rings  are  arranged  in  groups  and  the  pressure  to  each 
group  is  graded  by  suitable  connections  (Fig.  7). 

As  a  result  of  experiment  with  specially  constructed  apparatus 
it  was  found  the  regular  construction  would  not  answer  for  glands 
of  the  large  diameter  required  for  the  Carmania.  The  speeds  and 
pressures  were  so  high  that  the  rings  wore  rapidly.  Finally  radial 
fins,  as  in  Fig.  8,  were  used  in  connection  with  a  row  of  rings  and 
grooves.  The  action  of  these  fins  is  to  alternately  wire-draw  and 
expand  the  steam,  each  pair  constituting  an  expansion  stage,  thus 
reducing  its  pressure  as  it  travels  outward.  The  actual  gland  was 
fitted  at  each  end  of  both  the  high-  and  low-pressure  turbines,  as 
illustrated  in  Fig.  6,  with  four  rings,  K,  at  the  outer  end.  The 
small  amount  of  steam  which  is  allowed  to  leak  past  them  for  the 
purpose  of  lubrication  collects  in  pocket  G,  whence  it  is  led  by  the 
pipe  H  to  the  auxiliary  condenser  or  exhaust  tank.  In  the  case  of 
the  high-pressure  turbine,  where  the  radial  fins  do  not  sufficiently 
reduce  the  pressure  of  the  escaping  steam,  the  pocket  0  is  con- 
nected to  an  expansion  row  in  the  low-pressure  turbine. 

Comparison  with  the  Coronia. — The  area  occupied  by  the  tur- 
bines and  auxiliaries  is  practically  the  same  as  required  for  the 
quadruple-expansion  reciprocating  engines  of  the  sister  ship 
Coronia,  built  sometime  previously.  The  required  head  room  is 
less,  but  no  advantage  is  taken  of  this,  as  the  space  above  the  en- 
gine room  was  left  open  for  light  and  air.  There  is  a  saving  in 
weight  of  about  five  per  cent.  The  boiler  pressure  in  the  Coronia 
is  210  pounds  and  in  the  Carmania  195  pounds  per  square  inch. 
The  turbines  take  steam  at  an  initial  pressure  of  150  pounds  as 
against  200  pounds  in  the  quadruple  engines.  The  cooling  surface 
of  the  condensers  is  increased  in  the  Carmania  about  20  per  cent, 
the  capacity  of  the  centrifugal  pumps  is  about  double,  and  the 
weight  of  circulating  water  is  from  50  to  60  times  the  weight  of 
feed  water  as  compared  with  a  ratio  of  25  or  30  times  in  the 
Coronia's  installation. 

The  Lusitania  and  Mauritania,  of  the  Cunard  Line,  which  are 
expected  to  become  the  queens  of  the  sea,  are  turbine  vessels 
designed  to  maintain  a  minimum  speed  of  24  to  25  knots.  The 
dimensions  of  the  Lusitania,  which  is  more  nearly  completed  than 


STATUS  OF  THE  MARINE  TURBINE  403 

the  sister  ship,  are  a  length  of  785  feet  over  all,  extreme  breadth 
of  88  feet  and  a  depth  of  60>^  feet.  There  are  25  Scotch  boilers, 
all  but  two  of  which  are  double-ended,  and  carrying  a  pressure  of 
about  200  pounds  per  square  inch.  The  turbines  are  to  develop 
70,000  horse-power,  an  increase  of  70  per  cent  over  that  de- 
veloped by  the  largest  and  fastest  vessel  previously  constructed. 
The  turbines  are  divided  into  four  units,  each  of  which  has  its 
own  propeller  shaft  and  propeller.  The  two  outer  shafts  carry 
the  high-pressure  turbines  and  the  inner  shafts  the  low-pressure 
and  backing  turbines,  making  six  in  all,  four  for  forward  motion 
and  two  for  backing.  The  two  high-pressure  propellers  are 
located  about  80  feet  forward  of  the  low-pressure  propellers  and 
are  carried  by  long,  tapering  tubes,  which  support  them  at  quite 
a  distance  from  the  sides  of  the  hull. 

The  longest  blades  on  the  low-pressure  end  are  20  inches  in 
length  and  the  peripheral  speed  of  the  rotating  blades  ranges  from 
100  to  150  feet  per  second.  The  blades  are  bound  together  with 
brass  arid  copper  wire  soldered.  The  casings  and  blades  for  one 
of  the  low-pressure  turbines  weigh  450  tons  and  are  made  up  of 
six  sections.  The  rotors  of  the  high-pressure  turbines  are  about 
seven  feet  in  diameter  and  25  feet  long,  while  those  of  the  low- 
pressure  turbines  are  somewhat  larger.  It  is  stated  that  such 
accuracy  is  being  attained  in  the  construction  of  the  turbines  that 
it  is  expected  to  operate  them  successfully  with  a  clearance  be- 
tween the  blade  tips  and  the  casing,  or  the  rotor,  of  only  %0  mch. 

The  construction  of  these  great  vessels  was  induced  by  the 
success  of  the  recent  new  and  fast  German  vessels  of  the  Ham- 
burg-American line  in  wresting  away  the  supremacy  of  the  seas. 
The  fastest  of  these,  the  Kaiser  Wilhelm  II. ,  has  attained  a  speed 
of  23*^  knots.  As  a  matter  of  pride  it  was  desired  to  restore  to 
the  English  merchant  marine  the  distinction  of  having  the  "largest 
and  fastest"  afloat,  and  as  a  matter  of  safety  to  the  government 
the  Admiralty  wanted  more  fast  merchant  vessels  that  could  be 
drafted  into  the  service  in  time  of  war.  To  obtain  the  increase  of 
speed  from  the  23^  knots  of  the  fastest  German  vessel  to  the 
25  knots  called  for  in  the  turbine  ships,  a  tremendous  increase  in 
engine  power  was  necessary,  amounting  to  70  per  cent,  as  stated 


404  STEAM  TURBINES 

above,  and  it  is  doubtful  if  such  great  power  could  be  successfully 
generated  in  the  hold  of  a  ship  by  means  of  reciprocating  engines. 
Both  the  machinery  and  hull  of  the  Lusitania  are  from  the 
Clydebank  Works  of  Messrs.  John  Brown  &  Co.,  Ltd. 

Comparison  Between  Turbines  and  Reciprocating  Engines. 

The  best  opportunity  for  comparing  the  performance  of  vessels 
fitted  with  turbines  and  engines  has  been  afforded  by  the  Midland 
Railway  Company's  four  boats  of  the  Heysham  Line  of  Great 
Britain.  Of  these,  the  Londonderry  and  Manxman  have  turbines 
and  the  Antrim  and  Donegal  reciprocating  engines.  The  London- 
derry, Antrim  and  Donegal  have  the  following  dimensions: 
Length  330  feet,  breadth  42  feet,  depth  25  feet,  6  inches.  The 
Manxman  is  of  the  same  length  and  depth,  but  has  a  breadth  of 
43  feet.  The  turbine  boat  Londonderry  carries  150  pounds  boiler 
pressure  and  the  others  200  pounds  pressure.  The  engines  of  the 
Antrim  and  Donegal  are  of  the  triple-expansion  type,  differing 
only  in  details,  and  drive  a  single,  three-bladed  propeller.  The 
turbines  of  the  Manxman  were  designed  for  25  per  cent  more 
power  than  those  of  the  Londonderry,  but  are  of  similar  construc- 
tion and  drive  three  three-bladed  screws  after  the  usual  manner. 
All  the  boats  have  high-grade  condensing  apparatus,  but  the 
Manxman  has  in  addition  a  Parsons  vacuum  augmenter,  for  pro- 
ducing a  high  vacuum. 

Official  Trials  of  Heysham  Line  Boats. — The  results  of  the 
official  trials  showed  the  two  boats  with  reciprocating  engines  to 
be  on  a  par  in  economy  and  to  use  practically  the  same  amount  of 
feed  water  under  like  conditions.  At  speeds  of  19  to  20  knots, 
however,  which  is  the  working  speed  of  all  the  vessels  in  service, 
the  water  consumption  of  the  turbine  steamer  Londonderry  was 
8  per  cent  less  and  of  the  Manxman  14  per  cent  less  than  of  the 
Antrim  and  Donegal,  while  throughout  a  speed  range  from  1  to  20 
knots  the  turbine  boats  showed  superior  economy.  Speed  trials 
were  run  between  the  two  turbine  boats  and  the  Antrim  and  the 
Londonderry  proved  about  one  knot  faster  and  the  Manxman  from 
one  to  two  knots  faster  than  the  Antrim,  under  like  conditions. 


STATUS  OF  THE  MARINE  TURBINE 


405 


Results  based  on  the  Log  Books.* — Later,  comparisons  were  in- 
stituted between  the  turbine  steamers  and  those  with  reciprocating 
engines,  based  on  the  log  books  in  which  the  daily  records  were 
kept  while  the  boats  were  in  regular  service.  During  a  part  of  this 
comparative  period  the  Manxman  was  not  on  the  same  route  as 
the  other  vessels,  so  that  she  could  not  be  consistently  compared 
with  the  Antrim  and  Donegal  during  the  entire  time.  Also,  the 
high-pressure  turbine  of  the  Londonderry  was  partially  wrecked, 
owing  to  the  blades  of  the  rotating  drum  coming  in  contact  with 
the  stationary  blades,  so  that  the  records  from  this  steamer  were 
interrupted  for  three  months.  Valuable  comparative  figures  were 
secured,  however,  and  are  summarized  in  table  below.  The  regular 
route  was  between  Heysham  and  Belfast,  one  vessel  plying  each 
way  every  night  except  Sunday.  The  comparisons  are  made  in 
each  case  between  vessels  running  in  opposite  directions  on  the 
same  days  and  the  table  gives  the  weight  of  coal  each  vessel  con- 
sumed on  a  given  number  of  trips,  exclusive  of  that  burned  when 
in  port,  which  of  course  does  not  affect  the  performance  of  the 
propelling  machinery. 

TABLE  SHOWING  RESULTS  OBTAINED  BY  STEAMERS  RUNNING  SIMUL- 
TANEOUSLY, BUT  IN  OPPOSITE  DIRECTIONS. 


Reciprocating 
Engines. 

Turbines. 

Number  of  Trips  ..          

Antrim. 
48 
35.6 
19.7 

Donegal. 
42 
36 
19.2 

Antrim. 
29 
38.6 
19.5 

Donegal. 
39 
38.7 
19.3 

Londonderry. 
48 
35.3 
19.5 

Londonderry. 
42- 
36.9 
19.6 

Manxman. 
29 
38.6 
20.3 

Manxman. 
39 
40.2 
20.3 

Average  Coal  per  Trip  tons 

Average  Speed  in  knots  

Number  of  Trips  

Average  Coal  per  Trip  tons 

Average  Speed  in  knots  

Number  of  Trips  

Average  Coal  per  Trip  tons 

Number  of  Trips  

Average  Coal  per  Trip  tons 

Average  Speed  in  knots  

An  economy  in  the  turbine  steamers  is  the  small  amount  of  oil 
required,  only  five  gallons  being  used  per  trip,  and  the  dispensing 

*  Reported  in  London  Engineering. 


406  STEAM  TURBINES 

with  the  services  of  two  oilers  usually  required.  There  is  the  ab- 
sence of  vibration,  but  there  is  also  the  inferiority  in  maneuvering 
from  rest  in  narrow  waters.  Experiments  made  during  the  trial 
trips  of  these  turbine  boats  indicate  that  they  may  be  brought  to 
rest  from  full  speed  in  about  1^  minutes.  This  is  a  good  result, 
but  experience  has  shown  the  inadequacy  of  the  backing  power 
from  rest,  when  the  side  propellers  are  less  efficient. 

Advantages  and  Disadvantages  of  the  Marine  Turbine. — The 
chief  advantages,  on  the  evidence  already  given,  appear  to  be 
absence  of  vibration,  making  the  turbine  boat  a  much  pleasanter 
passenger  craft;  greater  speed  with  the  same  amount  of  coal  or 
less  coal  at  the  same  speed,  when  running  at  or  near  normal 
speeds;  ease  of  manipulation;  slight  saving  in  weight;  less  oil; 
less  attendance ;  less  liability  of  racing. 

The  chief  disadvantages  are  the  lack  of  holding  power  of  the 
small  screws ;  the  diminished  power  of  the  reversing  turbines ;  and 
the  poorer  economy  at  low  speeds. 

Cavitation. — The  size  of  the  propellers  on  turbine  vessels  is 
limited  by  trouble  experienced  through  cavitation.  When  the 
speed  of  a  propeller  blade  exceeds  a  certain  amount,  depending 
upon  the  type,  the  head  or  pressure  is  not  sufficient  to  keep  up 
the  supply  of  water  to  the  propeller  and  a  partial  vacuum  is  formed 
back  of  the  blade,  in  consequence  of  which  the  efficiency  of  the 
propeller  drops  off.  This  action  is  known  as  cavitation.  Marine 
turbines  are  made  larger  in  diameter  for  a  given  power  than  land 
turbines  used  to  drive  electric  generators,  and  in  this  way  the 
speed  of  rotation  is  reduced  somewhat.  The  speed  still  remains 
so  high,  however,  that  the  size  of  propeller  must  be  reduced  to 
avoid  cavitation ;  and  then,  to  secure  a  sufficiently  low  thrust  per 
square  inch  of  propeller  area  the  blades  are  made  wide.  A  wide- 
bladed  propeller  has  usually  been  considered  inefficient,  due, 
probably,  to  the  increased  friction  that  such  a  propeller  would 
have  if  of  the  usual  large  diameter.  This  objection  is  not  so 
serious  with  small  turbine  propellers,  however,  which  show  a 
reasonably  high  efficiency. 


APPENDIX 


Figs.  1  and  2  on  the  succeeding1  pages  show  the  kinetic 
energy  of  a  steam  jet  in  foot-pounds,  and  Figs.  3  and  4 
the  velocity  of  a  steam  jet  in  feet  per  second,  under  the 
assumption  that  the  steam  is  initially  dry  and  that  it 
attains  its  velocity  as  the  result  of  adiabatic  expansion. 
The  pressures  indicated  on  each  curve  are  initial  pres- 
sures and  the  figures  at  the  left  of  the  diagrams  are  final 
pressures.  The  values  at  the  bottoms  of  the  diagrams 
give  the  energy  or  velocity,  as  the  case  may  be,  of  dry, 
saturated  steam  expanding  from  a  given  initial  to  a  given 
final  pressure.  For  example,  steam  at  a  pressure  of  165 
pounds  absolute  and  discharging  from  a  nozzle  at  95 
pounds  absolute  would  acquire  a  velocity  of  about  1,500 
feet  per  second,  if  there  were  no  losses  of  energy ;  and 
the  jet  would  develop  approximately  35,000  foot-pounds 
of  energy  per  pound  of  steam  flowing.  If  one  pound 
discharged  per  minute,  the  jet  would  develop  a  little 
over  one  horse-power. 


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Fig.  1.     Energy  of  a  Steam  Jet  in  Foot-pounds,  when  the  Steam  Expands  Adiabatic- 
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batically  between  different  Initial  and   Final  Pressures. 


tit- 


412 


APPENDIX 


TABLE  OF  THE  PROPERTIES  OF  SATURATED  STEAM. 


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101.99 

70  0 

1043  0 

1113  1 

0.133 

1.985 

334.6 

0.00299 

5 

162  34 

130  7 

1000  8 

1131.5 

.235 

1.842 

73.22 

.01366 

10 

193  25 

161.9 

979  0 

1140  9 

.284 

1.781 

38.16 

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15 

213  03 

181  8 

965  1 

1146.9 

.314 

1.747 

26.15 

.03826 

20 

227.95 

196.9 

954  6 

1151.5 

.336 

1  722 

19.91 

.05023 

25 

240  04 

209  1 

946.0 

1155.1 

.354 

1.704 

16.13 

.06199 

30 

250  27 

219  4 

938  9 

1158.3 

.369 

1.689 

13.59 

.07360 

35 

259.19 

228  4 

932  6 

1161.0 

.381 

1.677 

11.75 

.08508 

40 

267.13 

236  4 

927.0 

1163  4 

.392 

1  666 

10.37 

.09644 

45 

274.29 

243  6 

922  0 

1165  6 

.402 

1.656 

9.287 

.1077 

50 

280.85 

250  2 

917.4 

1167  6 

.411 

1.648 

8  414 

.1188 

55 

286.89 

256  3 

913  1 

1169.4 

.419 

1.640 

7.696 

.1299 

60 

292.51 

261  9 

909  3 

1171.2 

.427 

1.634 

7.096 

.1409 

65 

297.77 

267  2 

905  5 

1172.7 

.434 

1.628 

6.583 

.1519 

70 

302.71 

272  2 

902  1 

1174.3 

.440 

1.622 

6.144 

.1628 

75 

307.38 

276.9 

898  8 

1175.7 

.446 

1.617 

5.762 

.1736 

80 

311.80 

281  4 

895.6 

1177.0 

.452 

1.612 

5.425 

.1843 

85 

316.02 

285  8 

892  5 

1178.3 

.458 

1.607 

5.125 

.1951 

90 

320  04 

290  0 

889  6 

1179.6 

.463 

1  603 

4.858 

.2058 

100 

327  58 

297  9 

884  0 

1181.9 

.473 

1.595 

4.403 

.2271 

115 

337.86 

308.7 

876.3 

1185.0 

.487 

1.584 

3.862 

.2589 

135 

350.03 

321.4 

867.3 

1188  7 

.503 

1.572 

3.323 

.3009 

140 

352.85 

324  4 

865.1 

1189.5 

.506 

1.570 

3.212 

.3113 

150 

358  26 

330.0 

861  2 

1191.2 

.513 

1  565 

3.011 

.3221 

165 

365.88 

338  0 

855.6 

1193.6 

.523 

1.558 

2.751 

.3635 

175 

370  65 

343  0 

852  0 

1195.0 

.529 

1  554 

2.603 

.3841 

185 

375  23 

347.2 

848.6 

1196.6 

.535 

1  550 

2.470 

.4049 

190 

377.44 

350  1 

847.0 

1197.1 

.538 

1  549 

2.408 

.4153 

200 

381  73 

354  6 

843.8 

1198  4 

543 

1  544 

2.294 

.4359 

215 

387.88 

361.0 

839.2 

1200  2 

.550 

1.539 

2.142 

.4669 

Abridged  from  the  tables  of  Prof.  C.  H.  Peabody,  with  his  permission. 
Values  of  the  entropy  of  steam  taken,  by  permission,  from  Heat  and 
Heat-engines,  by  Prof.  F.  R,  Hutton.  Both  of  the  foregoing  are  pub- 
lished by  John  Wiley  and  Sons,  New  York. 


INDEX 


Absolute    velocity 2 

calculation     of 286 

Accumulator,   Rateau's   steam 166 

calculations   for 169 

tests  on 171 

Adiabatic  flow ;  see  flow  of  steam. 

Air-pump,    Edwards 377 

Allis-Chalmers    turbine 153 

Angles   of   vanes.... 287,290-292,295 

experiments    on 298 

Area  of  condenser  surface 387 

floor,  for  engines  and  turbines..   332 

steam     nozzles 278 

Arrangement   of   condensers 337-341 

for    Curtis    turbine 337 

for  Parsons  turbine 339 

Balance     pistons 146,  155 

Balancing  high-speed  bodies 307 

cylinders    309 

locating   heavy   side 309 

static     308 

Blades,  see  vanes. 

Blowers,     turbine-driven • 79 

British- Westinghouse     turbine, 160 

Brown-Boveri   turbine 150 

Brownlee's    experiments   on    flow    of 

steam 218 

Buckets,  see  vanes. 

By-pass 143,  150,  152 

Care  and  management  of  turbines...   358 
condensing    apparatus. .  .360,  363,  369 

Carmania,   turbines  of 398 

compared    with    Coronia 400 

Cavitation 405 

Chart  for  power  units 193 

adiabatic  expansion : 

velocity  of  flow 408 

energy  of  flow 410 

condensation  during 270 

loss  of  superheat  during 275 

Commercial  aspect  of  the  turbine...    327 
Comparison   of  turbines  and  engines: 

advantages    of 329,  209 

calculations   for 176 

care     of 358 

cost     of 342 

economy    of 196,  329 

under  variable  loads 205,206 

with  overloads 209 

field    of 327,  328 

floor  area  for 332 


Comparison  of  turbines  and  engines: 

maintenance    of 345 

marine 393,  400,  403 

Composition  of  blades 357 

Compound  turbine,   principle  of....  16,21 

Condensing  surface,   area  of 387 

Cooling   water,   quantity  of 387 

Condensers  and  auxiliaries: 

Alberger    380 

arrangement    of 337-341 

for  Curtis  turbine 337 

for  Parsons  turbine 339 

calculations    on 389 

care   of 360,  363,  369 

cost    of 343 

injector  and  jet 385 

list  of  auxiliaries 375,380,382 

marine    400 

power    for 391 

space   required   for 337 

surface 375 

vs.    jet 384,  385 

underneath     turbine 339 

Wheeler   377 

Worthington     375,  379,  386 

Conversion   of  power  units 173,174 

Cost  of  engines,   turbinesj   etc 343 

maintenance  and   operation 345 

Critical     pressure 218,  277 

speed  of  rotating  bodies 305 

Curtis    turbine: 

care  and   operation  of 365-369 

condensers    for 337,  377,  378 

description    1 13-128 

governor    128 

patents    58-61 

principle    17,  58,  114 

rotative    speed 115 

sectional    view 115,  117 

small    sizes 124 

steam     pressures     and    velocities 

in 120 

stages   of 114 

step  bearing 122 

tests    on 188,  190,216 

valve  gear,  electric  type 126 

hydraulic     125 

mechanically  operated 125 

vanes,   construction 121 

diagram    for 293 


414 


INDEX 


Curtis  turbine: 

vertical     type 118 

De   Laval   turbine 67-79 

care     of 361 

gears  74 

governor    75 

nozzles   71,230 

oiling  arrangement 74 

patents  on 41,  47 

sectional    view 69 

speeds  of 70 

special  applications  of 77 

tests  on 183,  184 

Design,  notes  on  turbine 319 

example    in 325 

temperature-entropy  diagram  ap- 
plied   to 320 

of  steam  nozzles 276 

area  of 278 

diverging   280 

frictional   losses 281 

of  vanes 284-294 

Deterioration  of  turbines 352 

Dimensions   of   engines 336 

of    turbines 335 

Dow    turbine 46 

Dreadnought,  turbine  battleship 394 

Economy  of  engines,  average 194 

best  194 

in  commercial  operation 195 

in  street   railway  plants 209 

under  variable  loads 203 

with  varying  superheat 195 

Economy   of  small   engines   and   tur- 
bines         197 

Economy  of  turbines,  best 192 

miscellaneous   tables 183-192 

under  different  speeds 216 

under    heavy   overloads 210 

with  different  vacuums 211,212 

Efficiency      of      engine-type      gener- 
ators       177 

hydraulic    286 

steam  engines 178,  180 

steam  nozzles 220 

thermal  unit  basis  of 181 

turbines    314 

turbine    generators 177 

vanes    286,  292 

high    287,291 

Electric    governing 126 

Energy  of  a  jet 267 

see  charts  in  appendix. 

Enlargement   of  plant 341 

Entropy    254 

of  saturated  steam 256 

of  superheated  steam 257 

of  water ,.  255 


Erosion  of  blades 352 

cause  of 352 

in  De  Laval  turbines 353 

in  Parsons  turbines 353 

experience   with  Westinghouse 

turbines    355 

in  Curtis  turbines 356 

Experiments    with   nozzles ;   see   flow 
of  steam. 

upon  vanes 297 

upon  tubes  and  channels 302 

Floor     area     for     engines     and     tur- 
bines      332 

comparison  of  500  Kw.  units...   332 
Flow  of  steam : 

calculations  on 266-276 

saturated    266 

superheated     272 

condensation    during. . .  .260,  267,  270 

expansion    incomplete 269 

carried   too    far 225,279 

experiments     on 217-246 

Napier's  rules  for 217 

pressure    at    which    superheated 

steam  loses  superheat 273 

principles    of 9-11,  218,  219,  225 

shape  of  jets 9 

simplified  formula  for 269 

through        cylindrical        nozzles, 

Brownlee     219 

Kunhardt 219 

Kneass    224 

Gutermuth     240 

through       converging       nozzles, 

Rateau    236 

Gutermuth     240 

through      diverging      nozzles, 

Kneass   221 

Rosenhain    228 

Gutermuth     240 

Lucke   243 

through     orifice     in     flat     plate, 

Rateau    236 

Rosenhain    228 

weight  of  steam  flowing,  217,  238,  271 

Friction  of  steam  engines 179 

losses  in  nozzles 281 

Gears,  data  upon  De   Laval 74 

Generators,  efficiency  of 177 

encased    148,  330 

Glands,     water-packed 140 

Governors,  turbine: 

De   Laval 75 

Hamilton-Holzwarth     112 

Zoelly   107 

Curtis   128 

Greissmann's  tests  on  specific  heat..   263 
Grindley's  tests  on  specific  heat 263 


INDEX 


415 


Guide      vanes,      Hamilton-Holz- 

warth     109-112 

angles   of 287,  290-292,  295 

Rateau    96,  100 

Zoelly   104 

Hamilton-Holzwarth    turbine 109-112 

details  of  construction 110 

governor    112 

Heat  diagram 253 

applied   to   design 320 

Heat    unit 249 

Heat,  mechanical  equivalent  of 249 

specific   249 

to  laise  temperature  of  water...   251 

in  superheated  steam 252 

in   wet  steam 252 

latent    251 

total    251 

of    liquid 251 

High-speed    bodies 305 

critical   speed   of 305 

settling  of 307 

balancing 307 

stresses     in 310 

High  vacuum,  see  vacuum. 
Horse-power,    conversion    of   to   kilo- 
watts         173 

internal,  or  indicated 173 

conversion     table    for    kilowatts 

to     174 

Hydraulic  turbines,  principle  of 1 

governing    107,  125 

Impulse  turbines: 

and  reaction,  combined 159 

British- Westinghouse    160 

Crocker- Warren   159 

Sulzer    Bros 161 

compound : 

Hamilton-Holzwarth     109 

Kerr 93 

Rateau    95 

Zoelly   102 

distinguished  from  reaction 12 

shape  of  vanes 15 

simple : 

De   Laval 67-79 

Rateau   52,  81 

Riedler-Stumpf 81 

Injection  water,  see  cooling  water. 

lets,   impulse  and  reaction  of 3 

steam,  shape  of 9 

measuring  reaction  of 228 

Kerr    turbine 93 

Kilowatts,    conversion    of    to    horse- 
power       173 

King  Edward,  first  turbine  steamer..   392 
Kneass'     experiments     on     flow     of 

steam 221 


Knoblauch,  Linde  and  Klebe  tests  on 

specific   heat 263 

Kunhardt's    experiments    on    flow    of 

steam    220 

Latent    heat 251 

Leakage  of  steam 304 

in  engines 352 

Lindmark  turbine 162 

Losses  in  a  turbine,  analyzing 315 

in  1,250  Kw.  turbine 317 

Low-pressure  turbines 166 

Curtis   171 

steam  consumption  of 171 

Lubrication,  of  De  Laval  turbine...     74 

Parsons  turbine 141 

quantity  of  oil   required 330 

Lusitania,  largest  turbine  vessel 400 

Maintenance    and    operation 345 

labor    required 358 

Marine  turbines 392-405 

advantages  and  disadvantages...   405 

Atlantic  liners  with 394 

Carmania,  turbines  of 398 

compared    with    Coronia 400 

cavitation     405 

King  Edward,  turbines  of 392 

trials  of  Heysham  line  boats....   403 
of  Amethyst  and  Topaz 393 

Mechanical  equivalent  of  heat 249 

Motion,  absolute  and  relative 2 

Multicellular  turbines : 

explanation    of 20 

Hamilton-Holzwarth    109-1 12 

tests   on 187 

Rateau    95-102 

Zoelly    102-109 

Victorian,   turbines  of 394 

performance    of 396 

Napier's  rules  for  flow  of  steam 217 

National    Electric    Light   Association 

report   347 

Noise  made  by  turbines. 330 

Notation    247 

Nozzles : 

area  of 278 

converging  and  diverging.  .10,  11,  277 

cylindrical    10 

design  of 276 

effect  upon  steam  flowing 10 

of   maximum   efficiency 225,235 

taper  of  De  Laval 71,  230 

types  of 9 

see  flow  of  steam. 

Oiling  system,  De  Laval 74 

Parsons    141 

quantity  oil  required 330 

Operation  of  turbines: 

accumulator    system 369 


416 


INDEX 


Operation   of  turbines : 

condensing    apparatus. .  .360,  363,  367 

directions  by  engineers 364,  366 

general     care 362,  364,  365 

directions     359 

notes   of   experience 369 

oiling  system 368,  365,  362 

operating  Curtis  turbine ,. . .   365 

De  Laval  turbine 361 

Parsons     turbine 363 

synchronizing     368,  369 

warming  up 358,  359,  368 

Orifice    in    thin    plate,    effect    upon 

jet    12 

Orrok's  formulas  for  specific  heat...   263 
Overloads,  effect  on  turbine  economy,  210 

Parsons    turbines 135-158 

care    and   operation 359,  363,  364 

condenser  arrangements  for 

...339,  380,  382 

cost  of 343 

dimensions  of 335 

history    135 

patents    on 36,  42,  43,  54,  144 

principles     of 20,  136 

tests   on 185,  189,  190,210-212 

vanes,   diagrams  for 294 

Patents,  early  steam  turbine 22 

Altham  45 

Babbitt     41 

Bollmann     62 

Breguet     48 

Curtis    58-61 

Cutler     39 

De    Ferranti 55 

De    Laval 41,  47 

Delonchant    31 

Dow    46 

Foster  and  Avery 24 

Hartman    34 

Hoehl,  Brakell  and  Gunther. ...     36 

Imray   40 

Kerr    93 

Last    42 

Leroy    25 

Levin    91 

McElroy    51 

Monson    35 

Moorhouse    37 

Parsons    36,  42,  43,  54 

Perrigault  and  Farcot 36 

Pilbrow    27 

Rateau     52,  96 

Real  and  Pichon 23 

Richards    90 

Seger    49 

Tournaire    32 

Wilson    29 


Patents,   early   steam  turbine: 

Von   Rathen 57 

Zoelly 88 

Pelton  type,  turbines  of 80-94 

Kerr    turbine 93 

Levin's    experimental    wheel 92 

Rateau    81 

Richards'    design 90 

Riedler-Stumpf     81 

Zoelly     wheel 88 

Performance,  thermal  unit  basis  of..  181 
of     engines     and     turbines ;     see 
economy  of. 

Plant,  enlargement  of... 341 

Pressure,    atmospheric 248 

absolute    248 

critical   218,  277 

gauge    248 

in  throat  of  nozzles,  220,  221,  224,  227 

specific  250 

Pumps,     turbine-driven 77 

Reaction  of  jets,  how  measured 228 

Rateau  turbines : 

multicellular    95-102 

patents    on 52,  96 

simple  impulse 81 

test  on 187 

Reaction  turbines : 

distinguished  from  impulse 14 

shape  of  vanes 16 

Reduction  of  rotative  speed 16 

in    Curtis   turbine 115 

in  reaction  turbine :  21 

in    Riedler-Stumpf  system 84 

see  also  patents  of  Pilbrow,  Wil- 
son, Hartmann,  Moorhouse, 
Breguet  and  Ferranti,  Chap- 
ter II. 

Relative  velocity 2 

calculation  of 286 

Riedler-Stumpf  turbines : 

compound    129 

patents    on 83 

principle  of 17 

simple     impulse 81 

Rosenhain's  tests  on  flow  of  steam..  228 

Rotation  at  high  speed 305 

balancing   for 307 

stresses  caused   by 310 

Seger  turbine 49 

Space  for  condensing  apparatus 337 

Space   occupied  by   engines   and  tur- 
bines     332 

Speciric    heat 249 

of  superheated  steam 261 

see  tests  on 

Specific   pressure 250 

volume    . ,                                            .  249 


INDEX 


417 


Specific  pressure : 

of   superheated    steam 265 

of   wet  steam 252 

Speed  of  turbines,  effect  on  economy,  216 

Curtis    turbines 120,  124 

De   Laval  turbines 70 

see  reduction  of  rotative. 

Stage  turbines,  definition  of 19 

Curtis   114 

Steam,   flow  of 9,217-246,266-276 

generation    of 250 

saturated    250 

superheated     250 

specific  volume  of 265 

total    heat   of 252 

wet,  heat  in 251 

specific  volume  of 252 

Steam  accumulator  system  of  Rateau,  166 
Steam  engines: 

advantages    of 329 

compared    with     steam    turbines 
7,  8,  176,  196,  205,  206,  209,  329,  342 

cost  of 343 

dimensions    of 336 

field    of 328 

friction  tests  of 179 

leakage  in 352 

mechanical   efficiency  of 178,  180 

performance  of,  average 194 

best     194 

how  turbines  improve  upon...   209 

in  street  railway  plants 209 

under  variable   loads 204 

with    varying    superheat 195 

Steam  turbines: 

advantages    of 329 

cost     of 343 

commercial    aspect    of 327 

compared    with    steam     engines, 
7,  8,  176,  196,  205,  206,  209,  329,  342 

water  turbines 6 

compound   impulse 16,  95,  113 

compound     reaction 21,  135 

description  of 

A.    E.    G 132 

Allis-Chalmers    153 

Avery      24 

British-Westinghouse     160 

Brown-Boveri     150 

Curtis     17,58,113 

De    Laval 17,41,47,67 

Hamilton-Holzwarth     109 

Kerr    93 

Lindmark    162 

multicellular   95 

Parsons     20,  36,  42,  43,  54,  135 

Rateau     81,95 


Steam  turbines : 

Riedler-Stumpf     81,  129 

Seger    49 

stage    19,95,114 

Sulzer    Bros 160 

Westinghouse     137 

Zoelly     102 

deterioration  of 352 

dimensions  of 335 

electric  generating,  for 327,  330 

field     of 327 

limitations  of 327 

principles     of 1,  16 

simple    impulse 16,  67,  80 

troubles    of 346 

danger  from  water 349 

distortion  of  casing 349 

erosion    of  blades 352 

minor  difficulties 347 

stripping    the    blades 350 

Step  bearing  of  Curtis  turbine 123 

Stresses   in  rotating  bodies 310 

in  rotating  ring 311 

in    rotating    disk 312 

Sulzer   Brothers  turbine 161 

Superheated   steam 250 

engine  operating  with 195 

specific  heat  of 261 

specific  volume  of 265 

velocity  of  flow 273-275 

total   heat   of 252 

Surface  condensing  plants 375 

Taper  of  De  Laval  nozzles 71,230 

most   efficient   nozzles 225,235 

Temperature-entropy     diagram 253 

for  finding  condensation 260 

showing    reevaporation 322 

for    stage   turbine 320 

for  superheated  steam 259 

for  water  and  steam 257 

Temperature,  reference  points  of. ...   247 

absolute   248 

conversion    of    Fahr.    to    Cent . .   248 
of  steam  in  expanding  nozzle...   246 

Tests  on  channels  and  tubes 302 

condensers     388,  392 

generators   177 

nozzles;  see  flow  of  steam, 
specific      heat      of      superheated 
steam : 

Regnault    261 

Greissmann   263 

Grindley     263 

Knoblauch,    Linde  and    Klebe,  263 

turbine  boats 393,  400,  403 

vanes     297 

Tests  on  engines: 

average    results 194 


418 


INDEX 


Tests  on  engines: 

best     194 

friction   179 

in  commercial  operation 195 

in  street  railway  plants 209 

under  variable   loads 203 

with  varying  superheat 195 

Tests  on  turbines : 

at  different   speeds 216 

best     192 

under   heavy   overloads 210 

with  different  vacuums 211,212 

Curtis,  500  Kw.  at  Cork 188 

at    Newport 190 

2,000    Kw 188 

De  Laval,  at  different  loads....    183 

30   H.   P 183 

300   H.   P 184 

Parsons,     miscellaneous 185 

non-condensing      186 

Rateau     187 

Westinghouse-Parsons,   400   Kw.   189 

1,250    Kw 190 ' 

Zoelly     187 

Thermal  unit,  definition  of 249 

basis  of  performance 181 

Thrust,   end,   in   Parsons  turbine....    137 

in  multicellular  turbines 102 

i'otal    heat 251 

of  superheated  steam 252 

Turbine    vessels 392-405 

Car  mania    398 

Dreadnought     394 

King  Edward 392 

Lusitania     400 

Turbinia,    etc 392 

Vacuum : 

effect  on    engine  economy 371 

on    turbine   economy.  .372,  211,  212 

gain    from  high 373 

how    measured 248 

Vacuum  augmenter,  Parsons 382 

Valve,  secondary  admission. .  143,  150,  152 
Vanes : 

angles   of 287,  290-292,  295 

composition  of 357 

Curtis    turbine 121 

De  Laval  turbine 72 

diagrams     for 284 

compound  impulse  turbines...   283 

frictional    allowance 294 

high    efficiency 287 

Pelton    wheel 291 

reaction     turbines 294 

symmetrical     290 

erosion   of 352 

experimental     299 

guide,  see  guide  vanes. 


Vanes : 

impulse  and  reaction  upon 4 

Kerr's     93 

Rateau,    Pelton   type 52 

regular     type 97 

Richards'     design 88 

shape    of 15 

tests     upon 297 

Westinghouse     138 

Zoelly     89 

Variable  loads,  characteristics  under,  198 
steam-rate  curve  for  Parsons  tur- 
bine       199 

Rateau    turbine 200 

Curtis    turbine 202 

De  Laval  turbine 203 

Velocity  of   steam   flowing: 

calculation    of 266,  269 

for  superheated  steam 272-275 

experiments  on,  Kneass 223,227 

Rosenhain    233-235 

see  charts  in  appendix. 

Volume,    specific 249 

of   superheated    steam 265 

of  wet  steam 252 

Westinghouse-Parsons     turbines 137 

bearings   139 

care  and  operation  of. .  .359,  363,  364 
condenser  arrangement.  .339,  380,  382 

governing   arrangement 141 

indicator   diagram 142 

lubrication     141,  365 

secondary  admission  valve 143 

separate   high-   and  low-pressure 

cylinders    148 

sectional    view 146 

tests  on 189,  190,  210-212,  216 

water-packed  glands 140 

with    encased    generator 148 

vanes,  construction  of 138 

diagram    for 294 

Weight  of  steam  discharged: 

Napier's  rules  for 217 

calculation  of 71 

orifice  in  flat  plate 232,  239 

cylindrical    nozzles 220,  241 

converging  nozzles.  .232,  237,  241,  242 

diverging    nozzles 232,  241,  242 

Willans  and   Robinson  turbine 158 

Zoelly  turbine : 

tests  on 187 

patents  on 88 

Pelton  type,  vanes  of 89 

multicellular    102 

guides   and   vanes 104 

wheels  and  disks 107 

governor    107 

sectional    view 105 


AN    HUTIW- 

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